Control apparatus for compression-ignition type engine

ABSTRACT

The invention is provided with an ignition control section and an injection control section. When partial compression ignition combustion is carried out, the ignition control section causes an ignition plug to carry out: main ignition in which a spark is generated in a late period of a compression stroke or an initial period of an expansion stroke to initiate SI combustion; and preceding ignition in which the spark is generated at earlier timing than the main ignition. Also, when the partial compression ignition combustion is carried out, the injection control section causes an injector to inject fuel at such timing that the fuel exists in a cylinder at an earlier time point than the preceding ignition. Timing of the preceding ignition is more advanced when a swirl flow is gentle than when the swirl flow is strong.

TECHNICAL FIELD

The present invention relates to a control apparatus for acompression-ignition type engine capable of carrying out partialcompression ignition combustion to subject some of air-fuel mixture toSI combustion by spark ignition and to subject the rest of the air-fuelmixture to CI combustion by self-ignition.

BACKGROUND ART

Recently, attention has been paid to HCCI combustion. In the HCCIcombustion, gasoline fuel that is mixed with air is sufficientlycompressed in a combustion chamber and is burned by self-ignition. TheHCCI combustion is a mode of combustion in which air-fuel mixture issimultaneously and frequently burned. Thus, a combustion velocity of theair-fuel mixture is higher in the HCCI combustion than in spark ignitioncombustion (SI combustion) that is adopted in a normal gasoline engine,and it is said that the HCCI combustion is extremely advantageous overthe SI combustion in terms of thermal efficiency. However, the HCCIcombustion has a problem that combustion initiation timing (timing atwhich the air-fuel mixture is self-ignited) significantly fluctuates dueto an external factor such as a temperature, and also has a problem thatit is difficult to control the HCCI combustion during transientoperation that causes an abrupt change in an amount of a load.

For the above reason, it has been proposed to burn some of the air-fuelmixture by the spark ignition using an ignition plug instead of burningthe entire air-fuel mixture by self-ignition. That is, some of theair-fuel mixture is forcibly burned by flame propagation with the sparkignition as a start (SI combustion), and the rest of the air-fuelmixture is burned by the self-ignition (CI combustion). Hereinafter,such combustion will be referred to as partial compression ignitioncombustion.

As an example of an engine for which a similar concept to the partialcompression ignition combustion is adopted, an engine disclosed inPatent Literature 1 has been known. The engine disclosed in PatentLiterature 1 subjects stratified air-fuel mixture, which is producedaround the ignition plug (a spark plug) by auxiliary fuel injection, toflame propagation combustion by the spark ignition, and carries out mainfuel injection to the combustion chamber at a high temperature due toaction of the combustion (flame), so as to burn the fuel, which isinjected by this main fuel injection, by the self-ignition.

Meanwhile, it has also been proposed to improve the thermal efficiencyof the engine by another method not using compression ignitioncombustion. For example, Patent Literature 2 discloses a spark-ignitionengine that carries out the spark ignition by the ignition plug twiceper cycle. More specifically, in the spark-ignition engine disclosed inthe same literature, preceding ignition is carried out to supply suchlow ignition energy that not the entire air-fuel mixture in thecombustion chamber is ignited or burned (that a fire is locallygenerated) during a compression stroke, and at appropriate timing afterthis preceding ignition, main ignition is carried out to supply thehigher ignition energy than that in the preceding ignition. Just asdescribed, by producing the fire by the preceding ignition at an earlierstage than the main ignition, it is possible to prevent misfire of theair-fuel mixture and to increase the combustion velocity.

Here, in the partial compression ignition combustion, the combustionvelocity of the CI combustion has an impact on the thermal efficiency.Since the CI combustion is a phenomenon in which a spontaneous chemicalreaction of a fuel component occurs, it can be said that the combustionvelocity thereof is higher by nature than that of the SI combustion inwhich a combustion region is gradually expanded by the flamepropagation. Meanwhile, it is considered that, if a property of the fuelcan be modified to show high reactivity prior to the CI combustion, forexample, the combustion velocity of the CI combustion is furtherincreased, the thermal efficiency is thereby further improved, and thusfuel consumption performance and torque performance can be balanced.

The property of the fuel can possibly be modified to have the highreactivity when a temperature of the air-fuel mixture is increased tofall within a specified temperature range, for example. That is, whenthe fuel component (hydrocarbons) is cleaved due to the increase in thetemperature of the air-fuel mixture, an intermediate product includingOH radicals with the high reactivity is produced. As means forincreasing the temperature of the air-fuel mixture to modify theproperty of the fuel (to produce the intermediate product), theinventors of the present application considered to carry out pluraltimes of the spark ignition, that is, to carry out the auxiliarypreceding ignition prior to the main ignition so as to increase thetemperature of the air-fuel mixture as in the case of Patent Literature2 described above, for example. However, it was understood from thestudy by the inventors of the present application that, in the casewhere the preceding ignition was carried out at such timing that some ofthe air-fuel mixture was burned as in Patent Literature 2, a significantamount of the intermediate product was consumed by such combustion,which prevented an effect of increasing the combustion velocity of theCI combustion from being sufficiently exerted.

CITATION LIST Patent Literature

Patent Literature 1: JP-A-2009-108778

Patent Literature 2: Japanese Patent No. 4691373

SUMMARY OF INVENTION

The present invention has been made in view of circumstances asdescribed above and therefore has a purpose of providing a controlapparatus for a compression-ignition type engine capable of carrying outpartial compression ignition combustion at a high combustion velocityand with superior thermal efficiency.

In order to solve the problem, the present invention is a controlapparatus for controlling a compression-ignition type engine thatincludes: a cylinder; an injector that injects fuel into the cylinder;and an ignition plug that ignites air-fuel mixture, in which fuelinjected from injector and air are mixed, and that can carry out partialcompression ignition combustion to subject some of the air-fuel mixtureto SI combustion by spark ignition using the ignition plug and tosubject the rest of the air-fuel mixture to CI combustion byself-ignition. The control apparatus includes: a swirl generationsection that generates a swirl flow in the cylinder; an injectioncontrol section that controls fuel injection operation by the injector;and an ignition control section that controls ignition operation by theignition plug. When the partial compression ignition combustion iscarried out, the ignition control section causes the ignition plug tocarry out: main ignition in which a spark is generated in a late periodof a compression stroke or an initial period of an expansion stroke toinitiate the SI combustion; and preceding ignition in which the spark isgenerated at earlier timing than the main ignition. When the partialcompression ignition combustion is carried out, the injection controlsection causes the injector to inject the fuel at such timing that thefuel exists in the cylinder at an earlier time point than the precedingignition. Timing of the preceding ignition is set to be more advancedwhen the swirl flow generated by the swirl generation section is gentlethan when the swirl flow is strong.

With this configuration, it is possible to carry out partial compressionignition combustion at a high combustion velocity and with superiorthermal efficiency.

BRIEF DESCRIPTION OF DRAWINGS

FIG. 1 is a system view that schematically illustrates an overallconfiguration of a compression-ignition type engine according to anembodiment of the present invention.

FIG. 2 is a view illustrating a cross-sectional view of an engine bodyand a plan view of a piston.

FIG. 3 is a schematic plan view illustrating structures of a cylinderand intake/exhaust systems near the cylinder.

FIG. 4 is a block diagram illustrating an engine control system.

FIG. 5 is an operation map in which an engine operation range is dividedaccording to a difference in combustion mode.

FIG. 6 includes time charts for schematically illustrating combustioncontrol that is executed in each of the engine operation ranges.

FIG. 7 is a graph illustrating a waveform of a heat generation rateduring partial compression ignition combustion (SPCCI combustion).

FIG. 8 is a map illustrating a specific example of a target air-fuelratio that is set in a first operation range of the engine.

FIG. 9 is a map illustrating a specific example of a target swirlopening degree that is set in the first operation range.

FIG. 10 is a graph illustrating a change in the target swirl openingdegree in the case where a speed is changed under a condition that aload remains constant.

FIG. 11 is a view illustrating an outline of a rig tester that measuresa swirl ratio.

FIG. 12 is a graph illustrating a relationship between a swirl openingdegree and the swirl ratio.

FIG. 13 is a flowchart illustrating a specific example of combustioncontrol in a warm period of the engine.

FIG. 14 illustrates a subroutine in which details of control in step S10in FIG. 13 is illustrated.

FIG. 15 is a map for determining energy of preceding ignition.

FIG. 16 is a view for illustrating a characteristic of a calculationformula used to determine energy of preceding ignition.

FIG. 17 is a time chart illustrating, together with combustionwaveforms, an electric state of an ignition plug at the time when thepreceding ignition and the main ignition are carried out in the firstoperation range.

FIG. 18 is a graph illustrating a relationship between a temperature ofair-fuel mixture and a produced amount of an intermediate product.

FIG. 19 includes time charts illustrating a specific example of a casewhere the number of the preceding ignition is increased.

FIG. 20 is a graph illustrating a relationship between the number of thepreceding ignition and a fuel consumption improvement allowance.

FIG. 21 is a graph for illustrating various defining methods of an SIrate and corresponding to FIG. 7.

DESCRIPTION OF EMBODIMENTS

A description will hereinafter be made on an embodiment of the inventionwith reference to the accompanying drawings. The following embodiment isan example in which the present invention is embodied and does not havea property of limiting the technical scope of the present invention.

(1) OVERALL CONFIGURATION OF ENGINE

FIG. 1 and FIG. 2 are views illustrating a preferred embodiment of acompression-ignition type engine (hereinafter simply referred to as anengine), to which a control apparatus according to the present inventionis applied. The engine illustrated in the drawings is a four-cycledirect gasoline-injection engine that is mounted as a travel powersource on a vehicle, and includes: an engine body 1; an intake passage30 through which intake air to be introduced into the engine body 1flows; an exhaust passage 40 through which exhaust gas discharged fromthe engine body 1 flows; and an external EGR device 50 that circulatessome of the exhaust gas flowing through the exhaust passage 40 into theintake passage 30.

The engine body 1 has: a cylinder block 3 formed with a cylinder 2therein; a cylinder head 4 that is attached to an upper surface of thecylinder block 3 in a manner to close the cylinder 2 from above; and apiston 5 that is inserted in a slidingly reciprocal manner in thecylinder 2. Typically, the engine body 1 is of a multicylinder type thathas plural (for example, four) cylinders. However, a description willherein be made by focusing on the single cylinder 2 for simplificationof the description.

A combustion chamber 6 is defined above the piston 5, and thiscombustion chamber 6 is supplied with fuel having gasoline as a maincomponent by injection from an injector 15, which will be describedbelow. Then, the supplied fuel is mixed with air and burned in thecombustion chamber 6, and consequently, the piston 5 that is pushed downby an expansion force generated by the combustion reciprocatesvertically. The fuel that is injected into the combustion chamber 6 onlyneeds to contain gasoline as the main component and may contain asecondary component such as bioethanol in addition to gasoline, forexample.

A crankshaft 7 that is an output shaft of the engine body 1 is providedbelow the piston 5. The crankshaft 7 is coupled to the piston 5 via aconnecting rod 8, and is driven to rotate about a center axis accordingto reciprocal motion (vertical motion) of the piston 5.

A geometric compression ratio of the cylinder 2, that is, a ratiobetween a volume of the combustion chamber 6 at the time when the piston5 is at the top dead center and the volume of the combustion chamber atthe time when the piston 5 is at the bottom dead center is set to avalue that is suited for partial compression ignition combustion (SPCCIcombustion), which will be described below, and is equal to or higherthan 13 and equal to or lower than 30, preferably, equal to or higherthan 14 and equal to or lower than 18. In detail, the geometriccompression ratio of the cylinder 2 is preferably set to be equal to orhigher than 14 and equal to or lower than 17 in a case of a regularspecification for which gasoline fuel at octane rating of approximately91 is used, and is preferably set to be equal to or higher than 15 andequal to or lower than 18 in a case of a high-octane specification forwhich the gasoline fuel at the octane rating of approximately 96 isused.

The cylinder block 3 is provided with: a crank angle sensor SN1 thatdetects a rotation angle (a crank angle) of the crankshaft 7 and arotational speed of the crankshaft 7 (an engine speed); and a coolanttemperature sensor SN2 that detects a temperature of a coolant (anengine coolant temperature) flowing through the cylinder block 3 and thecylinder head 4.

The cylinder head 4 is provided with: an intake port 9 and an exhaustport 10, each of which is opened to the combustion chamber 6; an intakevalve 11 that opens/closes the intake port 9; and an exhaust valve 12that opens/closes the exhaust port 10. As illustrated in FIG. 2, a valvetype of the engine in this embodiment is a four-valve type having twointake valves and two exhaust valves. That is, the intake port 9includes a first intake port 9A and a second intake port 9B, and theexhaust port 10 includes a first exhaust port 10A and a second exhaustport 10B (see FIG. 3). The single intake valve 11 is provided for eachof the first intake port 9A and the second intake port 9B. The singleexhaust valve 12 is provided for each of the first exhaust port 10A andthe second exhaust port 10B.

As illustrated in FIG. 3, the second intake port 9B is provided with anopenable/closable swirl valve 18 (corresponding to an example of the“swirl generation section” or the “swirl adjustment device” in theclaims).

The swirl valve 18 is only provided in the second intake port 9B and isnot provided in the first intake port 9A. When such a swirl valve 18 isdriven to be closed, a ratio of the intake air that flows into thecombustion chamber 6 from the first intake port 9A, which is notprovided with the swirl valve 18, is increased. As a result, a spiralflow that circles around a cylinder axis Z (a center axis of thecombustion chamber 6), that is, a swirl flow can be intensified. On theother hand, when the swirl valve 18 is driven to be opened, the swirlflow can be weakened. The intake port 9 in this embodiment is a tumbleport that can generate a tumble flow (a vertical vortex). Thus, theswirl flow that is generated at the time when the swirl valve 18 isclosed is an oblique swirl flow that is mixed with the tumble flow.

The intake valves 11 and the exhaust valves 12 are driven to beopened/closed in an interlocking manner with the rotation of thecrankshaft 7 by valve mechanisms 13, 14 that include a pair of camshaftsand the like disposed in the cylinder head 4.

The valve mechanism 13 for the intake valves 11 includes an intake VVT13 a capable of changing open/close timing of the intake valves 11.Similarly, the valve mechanism 14 for the exhaust valves 12 includes anexhaust VVT 14 a capable of changing open/close timing of the exhaustvalves 12. The intake VVT 13 a (the exhaust VVT 14 a) is a so-calledphase-type variable mechanism and simultaneously changes the open timingand the close timing of the intake valves 11 (the exhaust valves 12) bythe same amount. With control by these intake VVT 13 a and exhaust VVT14 a, in this embodiment, it is possible to adjust a valve overlappingperiod in which both of the intake valve 11 and the exhaust valve 12 areopened across exhaust top dead center. In addition, it is possible toadjust an amount of burned gas (internal EGR gas) that remains in thecombustion chamber 6 by adjusting this valve overlapping period. Theintake VVT 13 a and the exhaust VVT 14 a correspond to examples of the“temperature adjustment device” in the claims.

The cylinder head 4 is provided with: an injector 15 that injects thefuel (mainly, the gasoline) into the combustion chamber 6; and anignition plug 16 that ignites air-fuel mixture produced by mixing thefuel injected into the combustion chamber 6 from the injector 15 and theair introduced into the combustion chamber 6. The cylinder head 4 isfurther provided with an in-cylinder pressure sensor SN3 that detects apressure in the combustion chamber 6 (hereinafter also referred to as anin-cylinder pressure).

As illustrated in FIG. 2, a cavity 20 is formed in a crown surface ofthe piston 5. The cavity 20 is formed by recessing a relatively largearea including a central portion of the crown surface in an opposite(downward) direction from the cylinder head 4. In addition, a squishedsection 21 having a ring-shaped flat surface is formed on a radiallyouter side of the cavity 20 in the crown surface of the piston 5.

The injector 15 is an injector of a multiple injection-port type thathas plural injection ports at a tip, and can radially inject the fuelfrom the plural injection ports (F in FIG. 2 represents a spray of thefuel injected from each of the injection ports). The injector 15 isarranged in a central portion of a ceiling surface of the combustionchamber 6 such that the tip thereof opposes the central portion of thecrown surface of the piston 5 (a center of a bottom portion of thecavity 20).

The ignition plug 16 is arranged at a position that is substantiallyshifted to the intake side from the injector 15. A position of a tip (anelectrode) of the ignition plug 16 is set to overlap with the cavity 20in a plan view.

As illustrated in FIG. 1, the intake passage 30 is connected to a sidesurface of the cylinder head 4 in a manner to communicate with theintake port 9. The air (fresh air) that is suctioned from an upstreamend of the intake passage 30 is introduced into the combustion chamber 6through the intake passage 30 and the intake port 9.

In the intake passage 30, an air cleaner 31, an openable/closablethrottle valve 32, a supercharger 33, an intercooler 35, and a surgetank 36 are sequentially provided from an upstream side. The air cleaner31 removes foreign substances in the intake air, the throttle valve 32adjusts a flow rate of the intake air, the supercharger 33 compressesthe intake air and delivers the compressed intake air, and theintercooler 35 cools the intake air that is compressed by thesupercharger 33.

An airflow sensor SN4 that detects the flow rate of the intake air,first and second intake temperature sensors SN5, SN7, each of whichdetects the temperature of the intake air, and first and second intakepressure sensors SN6, SN8, each of which detects a pressure of theintake air, are provided in portions of the intake passage 30. Theairflow sensor SN4 and the first intake temperature sensor SN5 areprovided in the portion of the intake passage 30 between the air cleaner31 and the throttle valve 32 and detects the flow rate and thetemperature of the intake air that flows through such a portion. Thefirst intake pressure sensor SN6 is provided in a portion of the intakepassage 30 between the throttle valve 32 and the supercharger 33 (on adownstream side of a connection port of an EGR passage 51, which will bedescribed below), and detects the pressure of the intake air that flowsthrough such a portion. The second intake temperature sensor SN7 isprovided in a portion of the intake passage 30 between the supercharger33 and the intercooler 35 and detects the temperature of the intake airthat flows through such a portion. The second intake pressure sensor SN8is provided in the surge tank 36 and detects the pressure of the intakeair in the surge tank 36.

The supercharger 33 is a mechanical supercharger that mechanicallycooperates with the engine body 1. A type of the supercharger 33 is notparticularly specified. For example, any of known Lysholm, root-type,and centrifugal superchargers can be used as the supercharger 33.

An electromagnetic clutch 34 capable of being electrically switchedbetween engagement and disengagement is interposed between thesupercharger 33 and the engine body 1. When the electromagnetic clutch34 is engaged, drive power is transmitted from the engine body 1 to thesupercharger 33, and the supercharger 33 supercharges the intake air.Meanwhile, when the electromagnetic clutch 34 is disengaged, thetransmission of the drive power is blocked, and the supercharger 33stops supercharging the intake air.

The intake passage 30 is provided with a bypass passage 38 for bypassingthe supercharger 33. The bypass passage 38 connects the surge tank 36and the EGR passage 51, which will be described below, with each other.The bypass passage 38 is provided with an openable/closable bypass valve39.

The exhaust passage 40 is connected to another side surface of thecylinder head 4 so as to communicate with the exhaust port 10. Theburned gas that is produced in the combustion chamber 6 is discharged tothe outside through the exhaust port 10 and the exhaust passage 40.

The exhaust passage 40 is provided with a catalytic converter 41. Thecatalytic converter 41 includes: a three-way catalyst 41 a for removingtoxic substances (HC, CO, NOx) contained in the exhaust gas flowingthrough the exhaust passage 40; and a gasoline particulate filter (GPF)41 b that catches particulate matters (PM) contained in the exhaust gas.On a downstream side of the catalytic converter 41, another catalyticconverter that includes an appropriate catalyst such as a three-waycatalyst or a NOx catalyst may be added.

A linear O₂ sensor that detects concentration of oxygen contained in theexhaust gas is provided in a portion of the exhaust passage 40 on anupstream side of the catalytic converter 41. The linear O₂ sensor is atype of a sensor, an output value of which is linearly changed accordingto a degree of the oxygen concentration, and an air-fuel ratio of theair-fuel mixture can be estimated on the basis of the output value ofthis linear O₂ sensor.

The external EGR device 50 has: the EGR passage 51 that connects theexhaust passage 40 and the intake passage 30; and an EGR cooler 52 andan EGR valve 53 provided in the EGR passage 51. The EGR passage 51connects a portion of the exhaust passage 40 on a downstream side of thecatalytic converter 41 and a portion of the intake passage 30 betweenthe throttle valve 32 and the supercharger 33. The EGR cooler 52 coolsthe exhaust gas (external EGR gas) that is recirculated from the exhaustpassage 40 into the intake passage 30 through the EGR passage 51 by heatexchange. The EGR valve 53 is provided to be openable/closable on adownstream side of the EGR cooler 52 (a side near the intake passage 30)in the EGR passage 51 and adjusts a flow rate of the exhaust gas flowingthrough the EGR passage 51. The EGR valve 53 corresponds to an exampleof the “temperature adjustment device” in the claims.

The EGR passage 51 is provided with a differential pressure sensor SN9that detects a difference between a pressure on an upstream side of theEGR valve 53 and a pressure on a downstream side thereof.

(2) CONTROL SYSTEM

FIG. 4 is a block diagram illustrating an engine control system. An ECU100 illustrated in the drawing is a microprocessor that integrallycontrols the engine, and includes an electric circuit including such asa CPU, a ROM, and a RAM, which are well-known. The ECU 100 correspondsto an example of the “controller” in the claims.

The ECU 100 receives detection signals from the various sensors. Forexample, the ECU 100 is electrically connected to the crank angle sensorSN1, the coolant temperature sensor SN2, the in-cylinder pressure sensorSN3, the airflow sensor SN4, the first and second intake temperaturesensors SN5, SN7, the first and second intake pressure sensors SN6, SN8,the differential pressure sensor SN9, and the linear O₂ sensor SN10described above. The ECU 100 sequentially receives information detectedby these sensors (that is, the crank angle, the engine speed, the enginecoolant temperature, the in-cylinder pressure, the flow rate of theintake air, the intake air temperature, the intake air pressure, thedifferential pressure between the upstream side and the downstream sideof the EGR valve 53, the oxygen concentration of the exhaust gas, andthe like).

The vehicle is also provided with an accelerator sensor SN11 thatdetects a pedal position of an accelerator pedal operated by a driverwho drives the vehicle, and the ECU 100 also receives a detection signalfrom this accelerator sensor SN11.

The ECU 100 makes various determinations and calculations on the basisof the input information from the above sensors while controlling eachcomponent of the engine. That is, the ECU 100 is electrically connectedto the intake VVT 13 a, the exhaust VVT 14 a, the injector 15, theignition plug 16, the swirl valve 18, the throttle valve 32, theelectromagnetic clutch 34, the bypass valve 39, the EGR valve 53, andthe like, and outputs a control signal to each of these devices on thebasis of a calculation result and the like.

More specifically, the ECU 100 functionally has a calculation section101, an injection control section 102, an ignition control section 103,a swirl control section 104, an intake control section 105, and an EGRcontrol section 106.

The injection control section 102 is a control module for controllingfuel injection operation by the injector 15. The ignition controlsection 103 is a control module for controlling ignition operation bythe ignition plug 16. The swirl control section 104 is a control modulefor controlling an opening degree of the swirl valve 18. The intakecontrol section 105 is a control module for adjusting the flow rate andthe pressure of the intake air to be introduced into the combustionchamber 6, and controls an opening degree of each of the throttle valve32 and the bypass valve 39 as well as ON/OFF of the electromagneticclutch 34. The EGR control section 106 is a control module for adjustingan amount of the EGR gas (the external EGR gas and the internal EGR gas)to be introduced into the combustion chamber 6, and controls operationof each of the intake VVT 13 a and the exhaust VVT 14 a as well as anopening degree of the EGR valve 53. The calculation section 101 is acontrol module for carrying out various calculations in order todetermine a control target value by each of these control sections 102to 106 and determine an operation state of the engine, for example. Thecalculation section 101 corresponds to an example of the “settingsection” or the “ignition timing setting section” in the claims, theignition control section 103 corresponds to an example of the“first/second ignition control sections” in the claims, and the EGRcontrol section 106 corresponds to an example of the “in-cylindertemperature adjustment section” in the claims.

(3) CONTROL ACCORDING TO OPERATION STATE

FIG. 5 is an operation map that is used in a warm period of the engine,and is a map illustrating differences in control according to thespeed/load of the engine. In the following description, that the engineload is high (low) is equivalent to that requested engine torque is high(low).

As illustrated in FIG. 5, when the engine is in a warm state, an engineoperation range is largely divided into three operation ranges A1 to A3according to a difference in combustion mode. When the operation rangesare set to a first operation range A1, a second operation range A2, anda third operation range A3, the third operation range A3 is a high-speedrange where the speed is high, the first operation range A1 is a low tomiddle-speed/low-load range where a portion on a high-load side iseliminated from a range on a low-speed side of the third operation rangeA3, and the second operation range A2 is a range that is the rest of theoperation ranges other than the first and third operation ranges A1, A3(in other words, a low to middle-speed/high-load range). Hereinafter, adescription will sequentially be made on the combustion mode selected ineach of the operation ranges, and the like.

(3-1) First Operation Range

In the low to middle-speed/low-load first operation range A1, thepartial compression ignition combustion (hereinafter this will bereferred to as SPCCI combustion) in which SI combustion and CIcombustion are combined is carried out. The SI combustion is acombustion mode in which the air-fuel mixture is ignited by a sparkgenerated from the ignition plug 16 and then the air-fuel mixture isforcibly burned by flame propagation that expands a combustion regionfrom the ignited point to a surrounding area. The CI combustion is acombustion mode in which the air-fuel mixture is self-ignited underhigh-temperature, high-pressure environment created by the compressionof the piston 5. The SPCCI combustion in which these SI combustion andCI combustion are combined is a combustion mode in which some of theair-fuel mixture in the combustion chamber 6 is subjected to the SIcombustion by the spark ignition that is carried out in environmentimmediately before the self-ignition of the air-fuel mixture and, afterthe SI combustion (due to the further increases in temperature and thepressure associated with the SI combustion), the rest of the air-fuelmixture in the combustion chamber 6 is subjected to the CI combustion bythe self-ignition. Here, “SPCCI” is an abbreviation for “sparkcontrolled compression ignition”.

The SPCCI combustion has such a property that the heat is more steeplygenerated in the CI combustion than in the SI combustion. For example,as illustrated in FIG. 6 and FIG. 7, which will be described below, in awaveform of a heat generation rate by the SPCCI combustion, a gradientat an initial rise at an initial stage of the combustion correspondingto the SI combustion is gentler than a subsequent gradient at an initialrise corresponding to the CI combustion. In other words, the waveform ofthe heat generation rate during the SPCCI combustion is created suchthat a first heat generation rate portion based on the SI combustion andhaving a relatively gentle gradient at an initial rise and a second heatgeneration portion based on the CI combustion and having a relativelysteep gradient at an initial rise occur consecutively in this order. Ina manner to correspond to such a tendency of the heat generation rate,in the SPCCI combustion, a pressure increasing rate (dp/dθ) in thecombustion chamber 6 that is generated in the SI combustion is lowerthan that in the CI combustion.

When the temperature and the pressure in the combustion chamber 6 areincreased by the SI combustion, the unburned air-fuel mixture isself-ignited in association therewith, and the CI combustion isinitiated. As exemplified in FIG. 6 and FIG. 7, which will be describedlater, the gradient of the waveform of the heat generation rate ischanged from being gentle to being steep at timing of this self-ignition(that is, timing at which the CI combustion is initiated). That is, thewaveform of the heat generation rate in the SPCCI combustion has aninflection point (X2 in FIG. 7) that appears at the initiation timing ofthe CI combustion.

After the initiation of the CI combustion, the SI combustion and the CIcombustion are carried out in parallel. A combustion velocity of theair-fuel mixture is higher in the CI combustion than in the SIcombustion. Thus, the heat generation rate is relatively high in the CIcombustion. However, the CI combustion is carried out after the piston 5reaches the compression top dead center. Thus, the gradient of thewaveform of the heat generation rate never becomes excessively steep.That is, after passing the compression top dead center, the piston 5lowers and reduces a motoring pressure. As a result, the increase in theheat generation rate is suppressed, which avoids an excessive increasein dp/dθ in the CI combustion. Just as described, in the SPCCIcombustion, due to a characteristic that the CI combustion is carriedout after the SI combustion, the excessive increase in dp/dθ, whichserves as an index of combustion noise, is unlikely to occur. Thus,compared to the simple CI combustion (a case where the entire fuel issubjected to the CI combustion), it is possible to suppress thecombustion noise.

With termination of the CI combustion, the SPCCI combustion is alsoterminated. The combustion velocity is higher in the CI combustion thanin the SI combustion. Thus, compared to the simple SI combustion (thecase where the entire fuel is subjected to the SI combustion),combustion termination timing can be advanced. In other words, in theSPCCI combustion, in an expansion stroke, it is possible to bring thecombustion termination timing closer to timing at which the piston 5reaches the compression top dead center. In this way, the fuelconsumption performance can be improved in the SPCCI combustion whencompared to the simple SI combustion.

As a specific mode of the SPCCI combustion as described above, in thefirst operation range A1, such control is executed that the spark isgenerated from the ignition plug 16 for plural times and the air-fuelmixture is subjected to the SPCCI combustion at the time of the lastspark ignition. In this embodiment, the number of the spark ignition isset to two. In order to carry out the SPCCI combustion with the two-timeignition, in the first operation range A1, the ECU 100 controls each ofthe engine components as follows. In the following description, as termsthat specify timing of the fuel injection and the spark ignition, termssuch as an “early period”, a “middle period”, and a “late period” of acertain stroke, and terms such as a “first half” and a “latter half” ofthe certain stroke will be used, and these terms are based on thefollowing preconditions. That is, in the present specification, in thecase where the stroke such as an intake stroke or a compression strokeis equally divided into three, periods are defined as the “earlyperiod”, the “middle period”, and the “late period” in time sequence.Thus, for example, (i) the early period, (ii) the middle period, and(iii) the late period of the compression stroke respectively indicateranges of (i) before compression top dead center (BTDC) 180 to 120° CA,(ii) BTDC 120 to 60° CA, and (iii) BTDC 60 to 0° CA. Similarly, in thepresent specification, in the case where the stroke such as the intakestroke or the compression stroke is equally divided into two, periodsare defined as the “first half” and the “latter half” in time sequence.Thus, for example, (iv) the first half and (v) the latter half of theintake stroke respectively indicate ranges of (iv) the BTDC 360 to 270°CA and (v) BTDC 270 to 180° CA.

During the operation in the first operation range A1, the ignition plug16 carries out preceding ignition and main ignition. In the precedingignition, the ignition plug 16 generates the spark at timing that issufficiently advanced from the compression top dead center. In the mainignition, the ignition plug 16 generates the spark at closer timing tothe compression top dead center than the preceding ignition. Thepreceding ignition is carried out in any of the early period or themiddle period of the compression stroke (BTDC 180 to 60° CA). The mainignition is carried out in a period from the late period of thecompression stroke to an initial period of the expansion stroke (BTDC 60to ATDC 60° CA).

For example, at an operation point P1 on the low-load side in the firstoperation range A1, as illustrated in Chart (a) in FIG. 6, the ignitionplug 16 carries out the preceding ignition in the early period of thecompression stroke and carries out the main ignition in the late periodof the compression stroke. Similarly, at an operation point P2 with thehigher load than the operation point P1, as illustrated in Chart (b) inFIG. 6, the ignition plug 16 carries out the preceding ignition in theearly period of the compression stroke and carries out the main ignitionin the late period of the compression stroke. However, the timing of thepreceding ignition at the operation point P2 on the high-load side isset to be advanced from the timing of the preceding ignition at theoperation point P1 on the low-load side. This is linked to timing ofsecond injection (the last fuel injection in one cycle), which will bedescribed below. That is, the timing of the preceding ignition is moreadvanced on the higher-load side in a manner to be linked to the timingof the second injection such that a crank angle period from the secondinjection to the preceding ignition is maintained to be substantiallyconstant.

In the preceding ignition carried out at the timing that issubstantially advanced from the compression top dead center as describedabove, the flame propagation of the air-fuel mixture does not occur.Although details will be described below, this preceding ignition iscarried out for a purpose of producing an intermediate productcontaining OH radicals. The intermediate product is produced when thetemperature of the air-fuel mixture around the spark (an arc) isincreased to a target temperature that is equal to or higher than 850 Kand lower than 1140 K and the fuel component (hydrocarbons) is therebycleaved. In addition, in order to reliably prevent the flamepropagation, energy of the preceding ignition is set to be lower thanenergy of the main ignition. Accordingly, even when such precedingignition is carried out, the air-fuel mixture is not substantiallyflamed, and the SI combustion is not initiated.

Meanwhile, the main ignition with the high energy that is carried out atthe timing relatively close to the compression top dead center causesthe flame propagation of the air-fuel mixture and causes the SIcombustion. When the SI combustion is initiated, the temperature and thepressure of the combustion chamber 6 are increased, which causes the CIcombustion. That is, the main ignition triggers the SPCCI combustion,some of the air-fuel mixture in the combustion chamber 6 is burned bythe flame propagation (the SI combustion), and the rest of the air-fuelmixture is burned by the self-ignition (the CI combustion).

The injector 15 injects the fuel to be injected in one cycle in pluralparts, and injects at least some of the fuel during the intake stroke.In this embodiment, the number of the fuel injection is set to two. Thatis, during the operation in the first operation range A1, in a specifiedearlier period than the above-described preceding ignition, the injector15 injects the fuel by dividing the fuel for the first injection and thesecond injection. For example, at the operation point P1 on the low-loadside in the first operation range A1, as illustrated in Chart (a) inFIG. 6, the injector 15 initiates the first injection in the first halfof the intake stroke, and initiates the second injection in the latterhalf of the intake stroke. Similarly, at the operation point P2 with thehigher load than the operation point P 1, as illustrated in Chart (b) inFIG. 6, the injector 15 initiates the first injection in the first halfof the intake stroke, and initiates the second injection in the latterhalf of the intake stroke. However, initiation timing of the secondinjection at the operation point P2 on the high-load side is set to beadvanced from initiation timing of the second injection at the operationpoint P1 on the low-load side. In other words, the timing of the secondinjection is advanced with an increase in the load in the firstoperation range A1.

An amount (a total amount) and a split ratio of the fuel that isinjected from the injector 15 by split injection as described above areset to be variable according to the requested engine torque. Morespecifically, the total amount of the fuel, that is, a sum of a fuelinjection amount by the first injection and a fuel injection amount bythe second injection is set to be increased on the higher-load side withthe increased requested torque. More specifically, the split ratiobetween the first/second injections, that is, (the fuel injection amountby the first injection):(the fuel injection amount by the secondinjection) is set to have the smaller ratio of the first injection onthe higher-load side. For example, the split ratio between the first andsecond injection is set to be changed from approximately 9:1 to 6:4 fromthe low-load side to the high-load side in the first operation range A1.

The opening degree of the throttle valve 32 is set to such an openingdegree that a larger amount of the air than an amount of the air at thestoichiometric air-fuel ratio is introduced into the combustion chamber6 through the intake passage 30. That is, the opening degree of thethrottle valve 32 is set to be relatively large so that the air-fuelratio (A/F) as a weight ratio between the air (the fresh air) introducedinto the combustion chamber 6 through the intake passage 30 and the fuelinjected into the combustion chamber 6 by the first and second injectionbecomes higher than the stoichiometric air-fuel ratio (14.7). Then, thelarger amount of the air than the amount of the air at thestoichiometric air-fuel ratio is introduced into the combustion chamber6 through the intake passage 30. Just as described, in this embodiment,during the operation in the first operation range A1, control to createsuch environment that the air-fuel ratio in the combustion chamber 6 ishigher than the stoichiometric air-fuel ratio (hereinafter this will bereferred to as lean A/F environment) while subjecting the air-fuelmixture to the SPCCI combustion is executed.

The air-fuel ratio (A/F) in the first operation range A1 is set to bevariable within a range that is higher than 20 and lower than 35. FIG. 8is a map illustrating a setting example of a target air-fuel ratio thatis a target value of the air-fuel ratio (A/F) in the first operationrange A1. As illustrated in this map, the target air-fuel ratio in thefirst operation range A1 is set to be substantially increased with theincrease in the load (the requested torque) in the first operation rangeA1. In detail, the target air-fuel ratio is set to have the highestvalue (equal to or higher than 31) in a range a1 that is set near anupper limit load in the first operation range A1 (that is, a load on aboundary between the first operation range A1 and the second operationrange A2) and to have a lower value as being separated from such a rangea1. However, the air-fuel ratio never becomes equal to or lower than 20at any position in the first operation range A1. In this embodiment, therange a1 with the maximum target air-fuel ratio is set to a belt-likerange that is separated from the upper limit load of the first operationrange A1 to the slightly low-load side and is separated from the lowerlimit speed of the first operation range A1 to the high-speed side, thatis, a middle to high-speed/high-load range in the first operation rangeA1. Since the range a1 is located near the upper limit load, an idlerange with the lowest speed and load is the farthest from the range a1in the first operation range A1. The target air-fuel ratio in this idlerange is the lowest.

In an inner range of a supercharging line T illustrated in FIG. 5, thesupercharger 33 is brought into an OFF state. In an outer range of thesupercharging line T, the supercharger 33 is brought into an ON state.In the inner range of the supercharging line T where the supercharger 33is brought into the OFF state, that is, in a low-speed portion of thefirst operation range A1, the electromagnetic clutch 34 is disengaged,and the supercharger 33 and the engine body 1 are decoupled. Inaddition, the bypass valve 39 is fully opened. In this way, thesupercharger 33 stops supercharging the intake air. Meanwhile, in theouter range of the supercharging line T where the supercharger 33 isbrought into the ON state, that is, in a high-speed portion of the firstoperation range A1, the electromagnetic clutch 34 is engaged, and thesupercharger 33 and the engine body 1 are coupled. In this way, thesupercharger 33 supercharges the intake air. At this time, the openingdegree of the bypass valve 39 is controlled such that the pressure (asupercharging pressure) in the surge tank 36, which is detected by thesecond intake pressure sensor SN8, matches a predetermined targetpressure per engine operation condition (conditions such as the speedand the load). For example, as the opening degree of the bypass valve 39is increased, the flow rate of the intake air that flows reversely tothe upstream side of the supercharger 33 through the bypass passage 38is increased. As a result, the pressure of the intake air that isintroduced into the surge tank 36, that is, the supercharging pressureis reduced. The bypass valve 39 adjusts a reverse flow rate of theintake air, just as described, and thereby controls the superchargingpressure to the target pressure.

In order to bring the temperature of the combustion chamber 6(hereinafter also referred to as the in-cylinder temperature) to atemperature suited for the SPCCI combustion, in many ranges in the firstoperation range A1, the intake VVT 13 a and the exhaust VVT 14 arespectively drive the intake valves 11 and the exhaust valves 12 atsuch timing that internal EGR for leaving the burned gas in thecombustion chamber 6 can be carried out. That is, the intake/exhaustVVTs 13 a, 14 a drive the valves 11, 12 to generate the valveoverlapping period in which both of the intake/exhaust valves 11, 12 areopened across the exhaust top dead center, and keep opening the exhaustvalves 12 until the piston 5 passes the exhaust top dead center (untilan initial period of the intake stroke). In this way, the burned gasreturns to the combustion chamber 6 from the exhaust port 10, and theinternal EGR thereby occurs. The valve overlapping period is adjusted toachieve the in-cylinder temperature that is suited to acquire thedesired waveform of the SPCCI combustion (a target SI rate and a targetθci, which will be described below), in other words, to introduce anamount of the internal EGR gas that is required to achieve such atemperature into the combustion chamber 6. An internal EGR rate that isacquired by adjusting the valve overlapping period just as described,that is, a proportion of the internal EGR gas to a total gas amount inthe combustion chamber 6 generally tends to be increased on thelower-load side in the first operation range A1.

The EGR valve 53 is opened in many ranges in the first operation rangeA1 so as to achieve the in-cylinder temperature suited for the SPCCIcombustion. That is, the EGR valve 53 is opened to cause external EGR inwhich the exhaust gas is recirculated into the combustion chamber 6through the EGR passage 51. Although details will be described below,the opening degree of the EGR valve 53 is adjusted to achieve thein-cylinder temperature that is suited to acquire the desired waveformof the SPCCI combustion (the target SI rate and the target θci, whichwill be described below), in other words, to introduce an amount of theexternal EGR gas that is required to achieve such a temperature into thecombustion chamber 6. An external EGR rate that is acquired by adjustingthe opening degree of the EGR valve 53 just as described, that is, aproportion of the external EGR gas to the total gas amount in thecombustion chamber 6 generally tends to be increased with the increasein the speed or the load in the first operation range A1.

The opening degree of the swirl valve 18 is set to a lower openingdegree than a half-opening degree (50%). When the opening degree of theswirl valve 18 is reduced just as described, a large portion of theintake air that is introduced into the combustion chamber 6 is theintake air from the first intake port 9A (the intake port on a sidewhere the swirl valve 18 is not provided), and the strong swirl flow isgenerated in the combustion chamber 6. This swirl flow is intensified inthe intake stroke, remains until the middle of the compression stroke,and promotes stratification of the fuel. That is, such a concentrationdifference occurs that the concentration of the fuel in a centralportion of the combustion chamber 6 is higher than that in an outer area(an outer circumferential portion) thereof. A detailed description onspecific setting of the opening degree of the swirl valve 18 will bemade in (4) below.

(3-2) Second Operation Range

In the low to middle-speed/high-load second operation range A2, suchcontrol is executed to subject the air-fuel mixture to the SPCCIcombustion by one time of the spark ignition. In other words, in thesecond operation range A2, the preceding ignition in the above-describedfirst operation range A1 is eliminated, and only the main ignition iscarried out. In order to carry out the SPCCI combustion with theone-time ignition, in the second operation range A2, each of thecomponents of the engine is controlled by the ECU 100 as follows.

The ignition plug 16 carries out the spark ignition once in the periodfrom the late period of the compression stroke to the initial period ofthe expansion stroke. For example, at an operation point P3 in thesecond operation range A2, as illustrated in Chart (c) in FIG. 6, theignition plug 16 carries out the spark ignition once in the late periodof the compression stroke. Then, this spark ignition triggers the SPCCIcombustion, some of the air-fuel mixture in the combustion chamber 6 isburned by the flame propagation (the SI combustion), and the rest of theair-fuel mixture is burned by the self-ignition (the CI combustion).

The injector 15 injects the fuel at least once in the intake stroke. Forexample, at the operation point P3 in the second operation range A2, asillustrated in Chart (c) in FIG. 6, the injector 15 carries out thesingle fuel injection during the intake stroke so as to supply an entireamount of the fuel to be injected in one cycle. In the range other thanthe operation point P3 (for example, at an operation point on thelower-load side than the operation point P3 in the second operationrange A2), the fuel may be injected in two parts in the intake stroke.

The opening degree of the throttle valve 32 is set to such an openingdegree that the air amount corresponding to the stoichiometric air-fuelratio is introduced into the combustion chamber 6 through the intakepassage 30, that is, the air-fuel ratio (A/F) as the weight ratiobetween the air (the fresh air) and the fuel in the combustion chamber 6substantially matches the stoichiometric air-fuel ratio (14.7).Meanwhile, as will be described below, in the second operation range A2,the EGR valve 53 is opened, and the external EGR gas is introduced intothe combustion chamber 6. Accordingly, in the second operation range A2,a gas air-fuel ratio (G/F) as a weight ratio between the whole gas andthe fuel in the combustion chamber 6 is higher than the stoichiometricair-fuel ratio (14.7). Just as described, in this embodiment, during theoperation in the second operation range A2, control to create suchenvironment that the gas air-fuel ratio (G/F) is higher than thestoichiometric air-fuel ratio and the air-fuel ratio (A/F) substantiallymatches the stoichiometric air-fuel ratio (hereinafter this will bereferred to as lean G/F environment) while subjecting the air-fuelmixture to the SPCCI combustion is executed.

The supercharger 33 is in the OFF state in a portion on the low-load andlow-speed side that overlaps the inner range of the supercharging lineT, and is in the ON state in the other ranges. When the supercharger 33is in the ON state and the intake air is supercharged, the openingdegree of the bypass valve 39 is controlled such that the pressure inthe surge tank 36 (the supercharging pressure) matches the targetpressure.

The intake VVT 13 a and the exhaust VVT 14 a respectively drive theintake valves 11 and the exhaust valves 12 at such timing that theinternal EGR is substantially stopped.

The EGR valve 53 is opened to the appropriate opening degree such thatan amount of the external EGR gas suited for the SPCCI combustion in thesecond operation range A2 is introduced into the combustion chamber 6.Similar to the case of the above-described first operation range A1, theopening degree of the EGR valve 53 at this time is adjusted to achievethe in-cylinder temperature that is suited to acquire the desiredwaveform of the SPCCI combustion (the target SI rate and the target θci,which will be described below).

The opening degree of the swirl valve 18 is set to a value that issubstantially equal to the opening degree thereof in the first operationrange A1 or to a specified intermediate opening degree that is largerthan the opening degree thereof in the first operation range A1.

(3-3) Third Operation Range

In the third operation range A3 on the high-speed side from the firstand second operation ranges A1, A2 described above, the relativelyorthodox SI combustion is carried out. In order to carry out this SIcombustion, in the third operation range A3, each of the components ofthe engine is controlled by the ECU 100 as follows.

The ignition plug 16 carries out the spark ignition once in the periodfrom the late period of the compression stroke to the initial period ofthe expansion stroke. For example, at an operation point P4 in the thirdoperation range A3, as illustrated in Chart (d) in FIG. 6, the ignitionplug 16 carries out the spark ignition once in the late period of thecompression stroke. Then, this spark ignition triggers the SIcombustion, and the entire air-fuel mixture in the combustion chamber 6is burned by the flame propagation.

The injector 15 injects the fuel for a specified period that at leastoverlaps the intake stroke. For example, at the operation point P4, asillustrated in Chart (d) in FIG. 6, the injector 15 injects the fuel fora succession of periods from the intake stroke to the compressionstroke. The quite high-speed and high-load conditions are set at theoperation point P4. Thus, the amount of the fuel to be injected in onecycle is large, and the crank angle period that is required to injectthe required amount of the fuel is extended. This is why the fuelinjection period at the operation point P4 is longer than the fuelinjection period at each of the other operation points (P1 to P3) thathas already been described.

The supercharger 33 is in the ON state, and the supercharger 33supercharges the intake air. The supercharging pressure at this time isadjusted by the bypass valve 39.

The opening degree of each of the throttle valve 32 and the EGR valve 53is controlled such that the air-fuel ratio (A/F) in the combustionchamber 6 becomes the stoichiometric air-fuel ratio or has a slightlyricher value (λ'1) than the stoichiometric air-fuel ratio.

The swirl valve 18 is fully opened. In this way, not only the firstintake port 9A but also the second intake port 9B is completely opened,and engine charging efficiency is improved.

(4) Swirl Control

Next, a detailed description will be made on swirl control in the firstoperation range A1. FIG. 9 is a map illustrating a specific example of atarget value of the opening degree of the swirl valve 18 (hereinafteralso referred to as a target swirl opening degree) that is set in thefirst operation range A1. FIG. 10 is a graph illustrating a change inthe target swirl opening degree in the case where the speed is changed(along a line V in FIG. 9) under a condition that the load remainsconstant. As illustrated in these drawings, in the first operation rangeA1, the target swirl opening degree is set to be variable within a rangeapproximately between 20 to 40%, and a value thereof is increased on thehigher-speed side or the higher-load side.

More specifically, in a first range b1 at the lowest speed and with thelowest load in the first operation range A1, the target swirl openingdegree is uniformly set to 20%. In a second range b2 at the higher speedor with the higher load than this first range b1, the target swirlopening degree is set to be gradually increased with the increase in thespeed or the load. In the second range b2, the target swirl openingdegree approximates 20% on the lower-speed/lower-load side near thefirst range b1, the target swirl opening degree becomes higher than 20%on the higher-speed/higher-load side away from the first range b1, andthe target swirl opening degree is increased up to approximately 40% atthe maximum. For example, in the case where the speed is increased in amanner to sequentially cross the first range b1→the second range b2(along the line V in FIG. 9), as illustrated in FIG. 10, the targetswirl opening degree is maintained at 20% while the speed is positionedin the first range b1. Thereafter, after the range is shifted to thesecond range b2, the target swirl opening degree is increased at asubstantially constant rate with the increase in the speed.

During the operation in the first operation range A1, the ECU 100 (theswirl control section 104) controls the opening degree of the swirlvalve 18 according to the maps (FIG. 9 and FIG. 10) for the target swirlopening degree that is set as described above.

As the opening degree of the swirl valve 18 is reduced, the strongerswirl flow is generated in the combustion chamber 6. In this embodimentin which the maps in FIG. 9 and FIG. 10 are used, during the operationin the first operation range A1, the opening degree of the swirl valve18 is reduced with the reductions in the speed and the load. Thus, theswirl flow is intensified according thereto (with the reductions in thespeed and the load). This is to promote the stratification of theair-fuel mixture and improve ignitability under a low-speed and low-loadconditions with difficulty in ignition.

That is, in this embodiment, the fuel is radially injected from theinjector 15, which is arranged in the center portion of the ceilingsurface of the combustion chamber 6, and each of the sprays of theinjected fuel is carried by the swirl flow and moves toward a centralportion of the combustion chamber 6. At this time, as the opening degreeof the swirl valve 18 is reduced (in other words, as an initial speed ofthe swirl flow is increased), the swirl flow remains until a late stageof the compression stroke. As a result, the air-fuel mixture with thehigh concentration of the fuel is produced in the central portion of thecombustion chamber 6 right up until the initiation of the combustion,which promotes the stratification of the air-fuel mixture. Based on sucha fact, in this embodiment, the opening degree of the swirl valve 18 isreduced to intensify the swirl flow under a condition of the low speedand the low load in the first operation range A1. In this way, theair-fuel mixture is stratified, and the ignitability thereof isimproved.

Here, the strength of the swirl flow will be defined. In the presentspecification, the strength of the swirl flow that is generated in thecombustion chamber 6 will be expressed as a “swirl ratio”. The swirlratio is defined as a value that is acquired by measuring a lateralangular velocity of an intake air flow per valve lifting, integratingthe measured values, and dividing the integral value by an angularvelocity of the crankshaft. The lateral angular velocity of the intakeair flow can be specified by measurement using a rig tester illustratedin FIG. 11. The rig tester illustrated in this drawing measures thelateral angular velocity of the intake air flow of a test engine thatincludes a cylinder block 203 and a cylinder head 204, and has: a basetable 210 arranged under the test engine; and an impulse meter 211arranged on the test engine. The test engine has a vertically flippedposture, and the cylinder head 204 thereof is placed on the base table210. An intake port 205 is formed in the cylinder head 204, and anintake supplier, which is not illustrated, is connected to this intakeport 205. A cylinder 202 is formed in the cylinder block 203, and theintake air that is supplied from the intake supplier is introduced intothe cylinder 202 via the intake port 205.

The impulse meter 211 has: a honey-comb rotor 211 a that is attached toan upper surface of the cylinder block 203; and a meter body 211 blocated on the honey-comb rotor 211 a. When a cylinder bore diameterthat is a diameter of the cylinder 202 is set as D, a lower surface ofthe impulse meter 211 is located at a position that is away from amating surface between the cylinder head 204 and the cylinder block 203by 1.75D. When the intake air is supplied from the intake supplier, theswirl flow (see an arrow in FIG. 11) is generated in the cylinder 202 inresponse thereto, and this swirl flow acts on the honey-comb rotor 211a. In this way, torque in a rotational direction is generated in thehoney-comb rotor 211 a. This torque is measured by the meter body 211 b,and the lateral angular velocity of the intake air flow is calculated onthe basis of the measured torque.

FIG. 12 illustrates a relationship between the opening degree of theswirl valve 18 in the engine of this embodiment and the swirl ratio thatis defined as described above. As illustrated in the drawing, the swirlratio is increased (that is, the swirl flow is intensified) with areduction in the opening degree of the swirl valve 18. For example, whenthe opening degree of the swirl valve 18 is 40%, the swirl ratio has avalue that slightly exceeds 1.5. Meanwhile, when the swirl valve 18 isfully closed (0%), the swirl ratio is increased to approximately 6.

Here, in this embodiment, as described above, during the operation inthe first operation range A1, the opening degree of the swirl valve 18is controlled within the range approximately between 20 to 40% (see FIG.9 and FIG. 10). It can be said from this that, in this embodiment, theopening degree of the swirl valve 18 in the first operation range A1 isset to such a value that the swirl ratio in the combustion chamber 6becomes equal to or higher than 1.5.

(5) REGARDING SI RATE

As described above, in this embodiment, the SPCCI combustion, in whichthe SI combustion and the CI combustion are combined, is carried out inthe first operation range A1 and the second operation range A2. In thisSPCCI combustion, it is important to control a ratio between the SIcombustion and the CI combustion according to the operating conditions.

Here, in this embodiment, as the above ratio, the SI rate that is aratio of a heat generation amount by the SI combustion to a total heatgeneration amount by the SPCCI combustion (the SI combustion and the CIcombustion) is used. FIG. 7 is a graph for illustrating this SI rate andillustrates a change in a heat generation rate (J/deg) by the crankangle at the time when the SPCCI combustion occurs. A point X1 on thewaveform in FIG. 7 is a heat generation point at which the heatgeneration rate rises in conjunction with the initiation of the SIcombustion, and a crank angle θsi corresponding to this heat generationpoint X1 is defined as initiation timing of the SI combustion. A pointX2 on the same waveform is an inflection point that appears when thecombustion mode is switched from the SI combustion to the CI combustion,and the crank angle θci corresponding to this inflection point X2 isdefined as initiation timing of the CI combustion. In addition, an areaR1 of the waveform of the heat generation rate that is located on anadvanced side (between θsi and θci) of θci as the initiation timing ofthis CI combustion is set as the heat generation amount by the SIcombustion, and an area R2 of the waveform of the heat generation ratethat is located on a retarded side of θci is set as the heat generationrate by the CI combustion. In this way, the above-described SI rate,which is defined by (the heat generation amount by the SIcombustion)/(the heat generation amount by the SPCCI combustion), can beexpressed as R1/(R1+R2) by using the areas R1, R2. That is, in thisembodiment, the SI rate=R1/(R1+R2).

In the CI combustion, the air-fuel mixture is simultaneously andfrequently burned by the self-ignition. Accordingly, the pressureincreasing rate tends to be higher in the CI combustion than in the SIcombustion by the flame propagation. Thus, in particular, in the casewhere the SI rate is reduced carelessly (that is, the ratio of the CIcombustion is increased) under a condition that the load is high and thefuel injection amount is large, large noise is produced. Meanwhile, theCI combustion does not occur until the temperature and the pressure ofthe combustion chamber 6 are sufficiently increased. Thus, under acondition that the load is low and the fuel injection amount is small,the CI combustion is not initiated until the SI combustion progresses tothe certain extent, and the SI rate is inevitably increased (that is,the ratio of the CI combustion is increased). In consideration of such acircumstance, in this embodiment, in the operation range where the SPCCIcombustion is carried out (that is, the first and second operationranges A1, A2), the target SI rate as a target value of the SI rate ispredetermined per engine operation condition. More specifically, in thefirst operation range A1 on the low-load side, the target SI rate is setto be generally reduced with the increase in the load (that is, theratio of the CI combustion is increased with the increase in the load).Meanwhile, the target SI rate in the second operation range A2 on thehigh-load side is set to be generally increased (that is, the ratio ofthe CI combustion is reduced) with the increase in the load.Furthermore, in association therewith, in this embodiment, the targetθci, which is the initiation timing of the CI combustion in the casewhere the combustion compatible with the target SI rate is carried out,is also predetermined per engine operation condition.

In order to obtain the target SI rate and the target θci describedabove, it is necessary to adjust control amounts such as the timing ofthe main ignition by the ignition plug 16, the injection amount/theinjection timing of the fuel from the injector 15, the EGR rates (theexternal EGR rate and the internal EGR rate) per operating condition.For example, as the timing of the main ignition is advanced, the largeramount of the fuel is burned by the SI combustion, which increases theSI rate. In addition, as the fuel injection timing is advanced, thelarger amount of the fuel is burned by the CI combustion, which reducesthe SI rate. Alternatively, as the in-cylinder temperature is increasedwith an increase in the EGR rate, the larger amount of the fuel isburned by the CI combustion, which reduces the SI rate. Furthermore, thechange in the SI rate is accompanied by a change in θci. Thus, a changein each of these control amounts (the main ignition timing, theinjection timing, the EGR rate, and the like) serves as an element usedfor adjustment of θci.

Based on the tendencies as described above, in this embodiment, the mainignition timing, the injection amount/the injection timing of the fuel,the EGR rate (and thus the in-cylinder temperature), and the like arecontrolled to make a combination capable of obtaining the target SI rateand the target θci described above at the time when the SPCCI combustionis carried out.

(6) CONTROL DURING SPCCI COMBUSTION

FIG. 13 is a flowchart illustrating details of combustion control(mainly control during the SPCCI combustion) that is executed in thewarm period of the engine. When the control illustrated in this drawingis started, in step S1, the calculation section 101 in the ECU 100calculates the requested engine torque on the basis of an acceleratoroperation state. That is, the requested torque as target torque to beoutput from the engine is calculated on the basis of an operation amount(a depression amount) and an operation speed of the accelerator pedalthat are specified from detection values of the accelerator sensor SN11.The higher requested torque is calculated as the operation amount andthe operation speed of the accelerator pedal are increased.

Next, in step S2, the calculation section 101 determines whether acurrent operation point of the engine is in the first operation range A1illustrated in FIG. 5. That is, the calculation section 101 specifiesthe operation point of the engine at a current time point on theoperation map in FIG. 5 on the basis of the engine speed detected by thecrank angle sensor SN1 and the requested torque calculated in step S1,and determines whether the current operation point is in the firstoperation range A1 of the map.

If it is determined NO in step S2 and thus is confirmed that the currentoperation point of the engine is not in the first operation range A1, instep S20, the calculation section 101 determines whether the currentoperation point is in the second operation range A2.

If it is determined YES in step S20 and is confirmed that the currentoperation point of the engine is in the second operation range A2, ascontrol corresponding to this second operation range A2, each of thecontrol sections 102 to 106 in the ECU 100 executes control (step S21)to subject the air-fuel mixture to the SPCCI combustion by the one-timespark ignition by the ignition plug 16. Contents of the control hereinare the same as that described above in (3-2). Thus, the detaileddescription thereon will not be made here.

If it is determined NO in step S20, that is, if it is confirmed that thecurrent operation point of the engine is in the third operation rangeA3, as control corresponding to this third operation range A3, each ofthe control sections 102 to 106 in the ECU 100 executes control (stepS22) to subject the air-fuel mixture to the SI combustion instead of theSPCCI combustion. Contents of the control herein are the same as thatdescribed above in (3-3). Thus, the detailed description thereon willnot be made here.

Next, a description will be made on control that is executed if it isdetermined YES in step S2, that is, if it is confirmed that the currentoperation point of the engine is in the first operation range A1. Inthis case, in step S3, the calculation section 101 in the ECU 100determines the target air-fuel ratio as the target value of the air-fuelratio (A/F) in the combustion chamber 6 on the basis of the requestedtorque (load) and the speed of the engine. That is, the calculationsection 101 determines the target air-fuel ratio that conforms to thecurrent operation point (the speed/the load) on the basis of therequested engine torque calculated in step S1, the engine speed detectedby the crank angle sensor SN1, and the map of the target air-fuel ratioillustrated in FIG. 8.

Next, in step S4, the calculation section 101 determines the injectionamount and the injection timing of the fuel to be injected from theinjector 15 on the basis of the requested engine torque calculated instep S1 and the engine speed detected by the crank angle sensor SN1. Theinjection amount/the injection timing of the fuel determined herein arethe predetermined injection amount/the injection timing per engineoperation condition to obtain the target SI rate and the target θcidescribed above. As illustrated in Charts (a) and (b) in FIG. 6, in thefirst operation range A1, the injection amount/the injection timing ofthe fuel are determined such that the fuel is injected by being dividedinto the first injection and the second injection and that the injectionamount is larger in the first injection than in the second injection.

In step S5, the calculation section 101 determines the opening degree ofthe throttle valve 32 on the basis of the target air-fuel ratiodetermined in step S3. That is, based on a precondition that the amountof the fuel determined in step S4 is supplied to the combustion chamber6, the calculation section 101 calculates such an opening degree of thethrottle valve 32 that an amount of the air (the fresh air)corresponding to the target air-fuel ratio is introduced into thiscombustion chamber 6, and determines this as a target opening degreevalue of the throttle valve 32.

Furthermore, in step S6, the calculation section 101 determines theopening degree of the swirl valve 18 on the basis of the requestedtorque (load) and the speed of the engine. That is, the calculationsection 101 specifies the opening degree of the swirl valve 18 thatconforms to the current operation point (the speed/the load) on thebasis of the requested engine torque calculated in step S1, the enginespeed detected by the crank angle sensor SN1, and the map of the swirlopening degree illustrated in FIG. 9, and determines this as a targetopening degree value of the swirl valve 18.

During the operation in the first operation range A1, in parallel withthe determination of the injection amount/the injection timing asdescribed so far, target control values related to the spark ignitionand the EGR (the external EGR/the internal EGR) are determined. That is,if it is determined YES in step S2 and is confirmed that the currentoperation point is in the first operation range A1, the processingproceeds to step S10, and the calculation section 101 determines thetiming and the energy of the preceding ignition by the ignition plug 16.A detailed description on this determination process will be made with acontrol flow in FIG. 14, which will be described below.

Next, in step S11, the calculation section 101 determines the target SIrate and the target θci on the basis of the requested engine torquecalculated in step S1. As described in above-described (5), the targetSI rate in the first operation range A1 is determined to be generallyreduced on the higher-load side where the requested torque is high (thatis, the ratio of the CI combustion is increased on the higher-loadside). In addition, the target θci is determined in association withthis determined target SI rate.

Next, in step S12, the calculation section 101 determines the timing ofthe main ignition by the ignition plug 16 on the basis of the target SIrate and the target θci determined in step S11. That is, the calculationsection 101 specifies: the initiation timing (θsi illustrated in FIG. 7)of the SI combustion that is required to achieve the combustionconforming to the target SI rate and the target θci; and the crank anglethat is advanced by a specified ignition delay time (a time requiredfrom the main ignition to the ignition) from the initiation timing θsiof this SI combustion, and determines this as a target value of the mainignition timing. The main ignition is normal spark ignition that iscarried out after a voltage of a capacitor included in an ignitioncircuit of the ignition plug 16 is increased to a maximum voltage. Thus,differing from the preceding ignition, it is not necessary to determinethe ignition energy according to the condition.

Next, in step S13, the calculation section 101 calculates thein-cylinder temperature that is required at a time point of the mainignition in order to obtain the target SI rate and the target θci, anddetermines this as a target in-cylinder temperature at the main ignitiontime point.

Next, in step S14, based on the target in-cylinder temperature at themain ignition time point calculated in step S13, the calculation section101 calculates the in-cylinder temperature that should be obtained atclose timing of the intake valve 11 (hereinafter also referred to asIVC) at which the compression of the combustion chamber 6 issubstantially initiated, that is, the target in-cylinder temperature atan IVC time point. This target in-cylinder temperature at the IVC timepoint is calculated on the basis of the target in-cylinder temperatureat the main ignition time point and an increase amount of thein-cylinder temperature that is estimated from a compression allowanceof the piston 5 from the IVC to the main ignition.

Next, in step S15, the calculation section 101 determines the openingdegree of the EGR valve 53 and the valve timing of the intake/exhaustvalves 11, 12 on the basis of the target in-cylinder temperature at theIVC time point calculated in step S14. That is, the calculation section101 calculates the external EGR rate and the internal EGR rate that arerequired to obtain the target in-cylinder temperature at the IVC timepoint on the basis of a difference between the target in-cylindertemperature at the IVC time point and the detected temperature by thefirst intake temperature sensor SN5 (that is, the temperature of thefresh air). Then, the calculation section 101 calculates the openingdegree of the EGR valve 53 that is required to obtain the calculatedexternal EGR rate, determines this as a target opening degree value ofthe EGR valve 53, calculates the valve timing of the intake/exhaustvalves 11, 12 required to obtain the calculated internal EGR rate, anddetermines this as target values of the valve timing.

Next, in step S16, the control sections (the injection control section102, the ignition control section 103, the swirl control section 104,the intake control section 105, and the EGR control section 106) in theECU 100 drive the injector 15, the ignition plug 16, the swirl valve 18,the throttle valve 32, the EGR valve 53, and the intake/exhaust VVTs 13a, 14 a on the basis of the various target control values determined inabove-described steps.

For example, the injection control section 102 controls the injector 15such that the amount of the fuel determined in step S4 is injected fromthe injector 15 at the determined timing.

As the preceding ignition, the ignition control section 103 controls theignition plug 16 such that the spark having the energy determined instep S10 is generated from the ignition plug 16. In addition, as themain ignition following this preceding ignition, the ignition controlsection 103 controls the ignition plug 16 such that the spark isgenerated from the ignition plug 16 at the timing determined in stepS12.

The swirl control section 104 controls the swirl valve 18 such that theopening degree of the swirl valve 18 matches the swirl opening degreedetermined in step S6.

The intake control section 105 controls the throttle valve 32 such thatthe opening degree of the throttle valve 32 matches the throttle openingdegree determined in step S5.

The EGR control section 106 controls the EGR valve 53 such that theopening degree of the EGR valve 53 matches the opening degree determinedin step S15, and controls the intake/exhaust VVTs 13 a, 14 a such thatthe intake/exhaust valves 11, 12 are opened/closed at timingcorresponding to the valve timing also determined in step S15.

With the control as described so far, in step S16, the air-fuel mixtureof the fuel injected into the combustion chamber 6 and the air issubjected to the preceding ignition and the main ignition, and isthereafter burned by the SPCCI combustion.

(7) DETERMINATION OF TIMING/ENERGY OF PRECEDING IGNITION

Next, a description will be made on a specific procedure at the timewhen the timing and the energy of the preceding ignition is determinedin above-described step S10. FIG. 14 illustrates a subroutine in whichdetails of the control in step S10 are illustrated. When this control isstarted, in step S31, the calculation section 101 in the ECU 100determines the energy of the preceding ignition. In this embodiment, theenergy of the preceding ignition is set in advance according to theengine operation condition, and the calculation section 101 sets theenergy of the preceding ignition to a value corresponding to the currentengine operation condition.

For example, the calculation section 101 stores the map illustrated inFIG. 15 and uses this map to determine the energy of the precedingignition. The map in FIG. 15 illustrates the relationship between anengine load and the energy of the preceding ignition in the case wherethe engine speed is set to be constant, in detail, in the case where thespeed is maintained to any of the low speed, the middle speed, and thehigh speed in the first operation range A1. As illustrated in FIG. 15,in this embodiment, the energy of the preceding ignition is specifiedfrom the speed and the load (the requested torque) of the engine and isgenerally set to be increased with the increase in the engine speed andthe load. In detail, under a condition that the engine speed isconstant, the energy of the preceding ignition is increased with theincrease in the load in a large portion of a load range except for arange where the engine load is extremely high (near the upper limit loadin the first operation range A1). However, in the range where the engineload is extremely high, reversely, the energy of the preceding ignitionis reduced with the increase in the load. Such a tendency also appearsin the case where the engine speed is maintained to any of the lowspeed, the middle speed, and the high speed.

In this embodiment, in step S31, the energy of the preceding ignition isdetermined by applying the engine load (the requested torque)/speed atthe current time point to the map in above FIG. 15. At this time, theenergy of the preceding ignition that is not defined in the map can becalculated by linear interpolation, for example. That is, FIG. 15defines a characteristic of the energy of the preceding ignition (therelationship between the engine load and the ignition energy) in thecase where the engine speed is any of the low speed, the middle speed,and the high speed. However, in the case where the engine speed is noneof the above three speeds, the energy of the preceding ignition can bedetermined by the linear interpolation using two prescribed values thatare close to each other. In order to improve accuracy of this linearinterpolation, a characteristic that corresponds to the different speedfrom the above three engine speeds (the low speed, the middle speed, andthe high speed) may be added to FIG. 15.

Next, in step S32, the calculation section 101 specifies swirl openingdegree, the engine coolant temperature, and the in-cylinder pressurethat are used as parameters for a calculation in step S34, which will bedescribed below. That is, at a time point at which timing (hereinafterreferred to as calculation timing) after a specified period from the IVCtime point as the close timing of the intake valve 11 arrives, thecalculation section 101 determines the swirl opening degree, which hasbeen determined by using the map in FIG. 9 in last step S6, as the swirlopening degree used for the calculation. The specified period is set inadvance and stored in the calculation section 101. Similarly, at thetime point at which the calculation timing arrives, the calculationsection 101 determines the engine coolant temperature, which has beendetected by the coolant temperature sensor SN2 at the most recenttiming, as the engine coolant temperature used for the calculation, anddetermines the in-cylinder pressure, which has been detected by thein-cylinder pressure sensor SN3 at the IVC time point, (that is, thein-cylinder pressure at the IVC time point) as the in-cylinder pressureused for the calculation.

Next, in step S33, the calculation section 101 estimates the air-fuelratio of the air-fuel mixture in the combustion chamber 6. In thisembodiment, during the operation in the first operation range A1, thefuel injection amount from the injector 15 and the opening degree of thethrottle valve 32 are controlled to values with which the targetair-fuel ratio illustrated in FIG. 8 can be obtained. However, theactual air-fuel ratio may vary from the target air-fuel ratio due to afluctuation in the air amount or the like. Thus, in order to grasp theaccurate air-fuel ratio near an actual value, the air-fuel ratio of theair-fuel mixture is estimated on the basis of various conditions. Morespecifically, the calculation section 101 calculates the amount of theair (the fresh air) that is actually introduced into the combustionchamber 6 on the basis of various parameters including the in-cylinderpressure at the IVC time point (the close timing of the intake valve 11)detected by the in-cylinder pressure sensor SN3, the flow rate of theintake air (the fresh air) detected by the airflow sensor SN4 before theIVC, the differential pressure between the upstream side and thedownstream side of the EGR valve 53 detected by the differentialpressure sensor SN9 before the IVC, and the valve timing of theintake/exhaust valves 11, 12 set by the intake/exhaust VVTs 13 a, 14 a.Then, the calculation section 101 calculates the air-fuel ratio (A/F) asa weight ratio between the air and the fuel on the basis of thecalculated air amount and the fuel injection amount from the injector 15determined in step S4, and determines this as an estimation value of theair-fuel ratio of the air-fuel mixture that is actually produced in thecombustion chamber 6.

In addition to estimation processing of the air-fuel ratio in step S33,the calculation section 101 constantly executes learning processingusing the linear O₂ sensor SN10. That is, the calculation section 101compares the air-fuel ratio estimated in step S33 with the air-fuelratio that is specified from the detection value of the linear O₂ sensorSN10, and, in the case where there is an error therebetween, corrects acalculation formula for the estimation in a manner to reduce the error.Such learning processing leads to improvement in estimation accuracy ofthe air-fuel ratio.

Next, in step S34, the calculation section 101 determines the timing ofthe preceding ignition from the specified calculation using each of theparameters (the swirl opening degree, the engine coolant temperature,the in-cylinder pressure, and the air-fuel ratio) specified in stepsS32, S33.

Here, a purpose of the preceding ignition is to modify a property of thefuel by applying the energy so as not to cause the flame propagation tothe air-fuel mixture that is not sufficiently compressed. In order toachieve this purpose, it is necessary to increase the temperature of theair-fuel mixture around the spark (the arc) by the preceding ignition tothe temperature that is equal to or higher than 850 K and lower than1140 K. By producing an air-fuel mixture layer in such a temperaturezone (hereinafter also referred to as a high-temperature portion), it ispossible to produce the intermediate product such as OH radicals bycleaving the fuel component (hydrocarbons) and to improve reactivitythereafter. The maps of the energy of the preceding ignition used instep S31 are predetermined to conform to such a purpose of the precedingignition, that is, to generate the high-temperature portion at equal toor higher than 850K and lower than 1140K around the spark.

However, likeliness of the occurrence of the flame propagation varies bythe environment (the temperature, the pressure, and the like) of thecombustion chamber 6. Accordingly, the timing of the preceding ignitionhas to be adjusted in consideration thereof. The calculation formulaused to determine the timing of the preceding ignition in step S34 ispredetermined in a manner to fulfill such a purpose, that is, such thata high-temperature portion at a temperature equal to or higher than 850K and lower than 1140 K is further reliably generated around the sparkeven when the environment in the combustion chamber 6 is changedvariously.

FIG. 16 is a view for illustrating a characteristic of the calculationformula used in step S34. As illustrated in FIG. 16, in step S34, thetiming of the preceding ignition is determined by using the specifiedcalculation formula that has, as input elements, the opening degree ofthe swirl valve 18 specified in step S32, the engine coolanttemperature, the in-cylinder pressure at the IVC time point, and theair-fuel ratio (A/F) of the air-fuel mixture in the combustion chamber 6estimated in step S33 and that has, as an output element, the timing ofthe preceding ignition. This calculation formula is set in advance in amanner to exhibit characteristics as illustrated in Graphs (a) to (d) inFIG. 15.

More specifically, the characteristics of the calculation formula arepresented graphically in Graphs (a) to (d) in FIG. 16. Graph (a)illustrates a relationship between the opening degree of the swirl valve18 and the timing of the preceding ignition, Graph (b) illustrates arelationship between the engine coolant temperature and the timing ofthe preceding ignition, Graph (c) illustrates a relationship between thein-cylinder pressure and the timing of the preceding ignition, and Graph(d) illustrates a relationship between the air-fuel ratio (A/F) and thetiming of the preceding ignition. The characteristics illustrated inthese Graphs (a) to (d) are those of a case where only a parameter on ahorizontal axis of each of the graphs is changed while the otherparameters are fixed.

For example, Graph (a) in FIG. 16 illustrates a changing tendency of thetiming of the preceding ignition in the case where only the openingdegree of the swirl valve 18 is changed while the parameters other thanthe opening degree of the swirl valve 18 are maintained to be constant.According to this Graph (a), the timing of the preceding ignition isadvanced with an increase in the opening degree of the swirl valve 18.In addition, as it is understood from an inclination of the graph, achange rate of the timing of the preceding ignition with respect to theopening degree of the swirl valve 18 is set to be lower in a range wherethe opening degree of the swirl valve 18 is small than in a range wherethe opening degree of the swirl valve 18 is large.

Graph (b) in FIG. 16 illustrates the changing tendency of the energy ofthe preceding ignition in the case where only the engine coolanttemperature is changed while the parameters other than the enginecoolant temperature are maintained to be constant. According to thisGraph (b), the timing of the preceding ignition is advanced with anincrease in the engine coolant temperature. In addition, as it isunderstood from an inclination of the graph, the change rate of thetiming of the preceding ignition with respect to the engine coolanttemperature is set to be lower in a range where the engine coolanttemperature is low than in a range where the engine coolant temperatureis high.

Graph (c) in FIG. 16 illustrates the changing tendency of the energy ofthe preceding ignition in the case where only the in-cylinder pressureis changed while the parameters other than the in-cylinder pressure aremaintained to be constant. According to this Graph (c), the timing ofthe preceding ignition is retarded with an increase in the in-cylinderpressure. In addition, as it is understood from an inclination of thegraph, the change rate of the timing of the preceding ignition withrespect to the in-cylinder pressure is set to be lower in a range wherethe in-cylinder pressure is high than in a range where the in-cylinderpressure is low.

Graph (d) in FIG. 16 illustrates the changing tendency of the timing ofthe preceding ignition in the case where only the air-fuel ratio ischanged while the parameters other than the air-fuel ratio aremaintained to be constant. According to this Graph (d), the timing ofthe preceding ignition is retarded with an increase in the air-fuelratio (as a ratio of the fuel becomes leaner). In addition, as it isunderstood from an inclination of the graph, the change rate of thetiming of the preceding ignition with respect to the air-fuel ratio isset to be lower in a range where the air-fuel ratio is high than in arange where the air-fuel ratio is low.

Here, a degree of an influence of each of the parameters, which are theopening degree of the swirl valve 18, the engine coolant temperature,the in-cylinder pressure, and the air-fuel ratio, on a flame propagationproperty mutually differs. Thus, the calculation formula is weighed inadvance in consideration of such a difference in the degree of theinfluence (sensitivity). For example, if the opening degree of the swirlvalve 18 has the largest influence on the flame propagation property,the calculation formula is weighed such that the tendency of Graph (a)in FIG. 16 is reflected most thereto.

As it has already been described, the timing of the preceding ignition,which is determined by using the calculation formula in FIG. 6 set tohave the characteristics as described so far, is set as such timing thatthe energy is lower than the energy of the main ignition and theproperty of the fuel is modified (the OH radicals and the like areproduced), in detail, such timing that the high-temperature portion atequal to or higher than 850 K and lower than 1140 K is generated aroundthe spark (the arc) by the preceding ignition.

(8) Specific Operation of Preceding Ignition/Main Ignition

As described above, in this embodiment, during the operation in thefirst operation range A1, the preceding ignition with the low energy andthe main ignition with the high energy are carried out in one cycle. Inorder to cause the ignition plug 16 to carry out such spark ignitionwith the different energy twice (the preceding ignition and the mainignition), the ignition plug 16 is controlled as follows, for example.

In this embodiment, the single ignition plug 16 is provided for thesingle cylinder 2, and the single ignition plug 16 includes the singleignition circuit constructed of an LC circuit including a coil, thecapacitor, and the like. Thus, in order to cause the ignition plug 16 tocarry out the spark ignition twice, it is required to repeatcharging/discharging of the capacitor.

FIG. 17 includes time charts illustrating an electric state of theignition plug 16 at the time when the preceding ignition and the mainignition are carried out in the first operation range A1 together with acombustion waveform, Chart (a) illustrates a waveform of the heatgeneration rate by the SPCCI combustion, Chart (b) illustrates awaveform of an energization command to the ignition plug 16, and Chart(c) illustrates a waveform of a discharging current from the ignitionplug 16. As indicated by waveforms W1, W2 in Chart (b) illustrated inFIG. 17, the ignition plug 16 is energized prior to the precedingignition and the main ignition. When energized times (so-called dwelltimes) are compared, the energization time (the waveform W1) at the timeof the preceding ignition is shorter than that (the waveform W2) at thetime of the main ignition. In addition, as indicated by waveforms Y1, Y2in Chart (c) of FIG. 17, the ignition plug 16 starts being discharged(the spark is generated) at a time point at which the energization ofthe ignition plug 16 is stopped. At this time, since the energizationtime for the preceding ignition is shorter than the energization timefor the main ignition, discharge energy (the waveform Y1) at the time ofthe preceding ignition is lower than the discharge energy (the waveformY2) at the time of the main ignition. This can also be understood froman area of the waveform Y1 being smaller than an area of the waveformY2.

In the example in FIG. 17, the energy that is stored by the energizationfor the preceding ignition (the waveform W1) is completely discharged bythe preceding ignition. This means that a voltage of the capacitor afterthe preceding ignition is substantially dropped to zero. Accordingly, inorder to store the sufficient energy for the main ignition in theignition plug 16, at the time of the energization for the main ignition(the waveform W2), it is necessary to continue the energization for arelatively long period so as to increase the voltage of the capacitorfrom zero to the maximum voltage. Meanwhile, in the preceding ignition,for which only the low energy is required, the energization can bestopped before the voltage of the capacitor reaches the maximum voltage.This is a reason why the energization time for the preceding ignition isshorter than the energization time for the main ignition.

Here, at least in the preceding ignition, it is not necessary todischarge the entire energy that has been stored before, and only someof the stored energy may be discharged. That is, in the case where theenergization of the ignition plug 16 is restarted while the ignitionplug 16 is discharged, the ignition plug 16 stops being discharged atthe time point. Thus, a larger amount of the energy than an amount ofthe energy that is originally required may be stored by theenergization, and then the energization may be restarted in the middleof discharging (in this way, discharging is stopped). In this way, onlysome of the stored energy may be released from the ignition plug 16. Inthe case where the preceding ignition by this method is carried out, theenergization time for the main ignition can be shortened. Thus, this isadvantageous in a case where an interval between the preceding ignitionand the main ignition is relatively short.

(9) OPERATIONAL EFFECTS

As it has been described so far, in this embodiment, in the firstoperation range A1 where the SPCCI combustion is carried out, the mainignition and the preceding ignition are carried out. In the mainignition, the spark is generated in the late period of the compressionstroke or the initial period of the expansion stroke to initiate the SIcombustion. In the preceding ignition, the spark is generated at thetiming that is earlier than that in the main ignition and is later thanthe fuel injection timing. Such a configuration has an advantage thatthe SPCCI combustion at the high combustion velocity and with thesuperior thermal efficiency can be carried out.

That is, in the above embodiment, the preceding ignition is carried outat the earlier timing than the main ignition, and the high-temperatureportion at equal to or higher than 850K and lower than 1140K isgenerated around the spark (the arc) by this preceding ignition.Accordingly, the property of the fuel can be modified such that theflame propagation of the air-fuel mixture does not occur by thepreceding ignition and that the thermal efficiency in the CI combustionis improved. In detail, the fuel component (hydrocarbons) is cleaved byheating to the above temperature zone so as to produce hydrogen peroxide(H₂O₂) and formaldehyde (CH₂O), and OH radicals from these componentscan be produced. OH radicals have strong oxidation behavior and are highin reactivity. Thus, when the intermediate product containing such OHradicals is produced in the combustion chamber 6 after the precedingignition, the combustion velocity of the CI combustion as the phenomenonin which the spontaneous chemical reaction of the fuel component occurscan be accelerated, which can improve the thermal efficiency.

FIG. 18 is a graph illustrating a relationship between the temperatureof the air-fuel mixture and a produced amount of the intermediateproduct, the graph being obtained by a numerical simulation performed bythe inventors of the present application. As indicated in this graph,the produced amount of the intermediate product is generally increasedwith the increase in the temperature of the air-fuel mixture. Athreshold α on a vertical axis of the graph represents the producedamount of the intermediate product that is required to exert asignificant effect, and indicates that, in the case where theintermediate organism, the amount of which is equal to or larger thanthe threshold α, exists in the combustion chamber, the combustionvelocity shows a significant difference. From the graph, in order toobtain the amount of the intermediate product that is equal to or largerthan the threshold α (that is, in order to accelerate the combustion ina significant level), it is necessary to increase the temperature of theair-fuel mixture at least to 850 K. The amount of the intermediateproduct keeps being increased even after the temperature of the air-fuelmixture becomes higher than 850 K. However, once the temperature reaches1140 K, the amount of the intermediate product is rapidly (almostvertically) reduced. It is considered that this is because the air-fuelmixture is burned, the flame is produced (that is, a flame reactionoccurs), almost all of the intermediate product is consumed when thetemperature of the air-fuel mixture reaches 1140K.

Meanwhile, in the above embodiment, the energy and the timing of thepreceding ignition are adjusted to such energy and such timing that thehigh-temperature portion at equal to or higher than 850 K and lower than1140 K is generated around the spark (the arc). Accordingly, it ispossible to reliably produce the intermediate product that contains OHradicals with the high reactivity by this preceding ignition, and it ispossible to improve the thermal efficiency by accelerating thecombustion velocity in the CI combustion. In Chart (a) of FIG. 17, thewaveform of the heat generation rate (a solid line) of a case where thepreceding ignition, for which the energy and the timing areappropriately adjusted just as described, is executed is compared to thewaveform of the heat generation rate (a broken line) of a case where thepreceding ignition is not executed. As apparent from the comparison ofthese two combustion waveforms, it is understood that the heatgeneration rate after initiation timing of the CI combustion (a retardedside from a point X′) rises rapidly and the combustion velocity in theCI combustion is increased when the preceding ignition is carried out incomparison with the case where the preceding ignition is not carriedout. The intermediate product that is produced by the preceding ignitionis partially consumed by the SI combustion prior to the CI combustion.At the initiation time point of the SI combustion, the intermediateproduct spreads over a large area of the combustion chamber 6, and theintermediate product remains on an outer side of the range of the SIcombustion. Thus, due to action of this remaining intermediate product,the CI combustion is accelerated with no difficulty.

Here, in order to modify the property of the fuel by the precedingignition as described above (to produce the intermediate product such asOH radicals without causing the flame propagation), it is necessary toappropriately adjust the timing of the preceding ignition according tovarious types of the environment in the cylinder 2. As a result of theearnest study on this point, the inventors of the present applicationhas obtained the following knowledge.

When the swirl ratio that represents the strength of the swirl flow ischanged, in detail, when only the swirl ratio is changed while theparameters other than the swirl ratio are maintained to be constant, thelimit timing of the ignition timing at which the flame propagationoccurs is retarded with an increase in the swirl ratio.

When the in-cylinder temperature is changed, in detail, when only thein-cylinder temperature is changed while the parameters other than thein-cylinder temperature are maintained to be constant, the limit timingof the ignition timing at which the flame propagation occurs is retardedwith the reduction in the in-cylinder temperature.

When the in-cylinder pressure is changed, in detail, when only thein-cylinder pressure is changed while the parameters other than thein-cylinder pressure are maintained to be constant, the limit timing ofthe ignition timing at which the flame propagation occurs is retardedwith the increase in the in-cylinder pressure.

When the concentration of the fuel in the combustion chamber 6 ischanged, in detail, when only the concentration of the fuel is changedwhile the parameters other than the concentration of the fuel aremaintained to be constant, the limit timing of the ignition timing atwhich the flame propagation occurs is retarded with the reduction in theconcentration of the fuel.

The method for determining the timing of the preceding ignition (FIG. 14and FIG. 16) that has been described in the above embodiment was devisedon the basis of the knowledge as described above. According to thismethod, the property of the fuel can reliably be modified while theoccurrence of the flame propagation by the preceding ignition isavoided.

That is, in the above embodiment, with the condition that, as theopening degree of the swirl valve 18 is increased, in other words, theswirl ratio is reduced and the swirl flow is weakened, the timing of thepreceding ignition is advanced (see Graph (a) in FIG. 16). Thus, it ispossible to apply the energy to the air-fuel mixture by the precedingignition at the appropriate timing that corresponds to the difference inthe flame propagation property by advancing the ignition timing when theswirl flow is gentle and the flame propagation is thereby likely tooccur or by retarding the ignition timing when the swirl flow is strongand the flame propagation is thereby less likely to occur. As a result,within the range where the flame propagation of the air-fuel mixturedoes not occur, it is possible to generate the high-temperature portionin such an appropriate temperature zone that the property of theair-fuel mixture is sufficiently modified (that is, the high-temperatureportion at equal to or higher than 850 K and lower than 1140 K) aroundthe spark, and it is thus possible to reliably produce the intermediateproduct that contributes to the acceleration of the CI combustion.

For the similar purpose, in the embodiment, as the engine coolanttemperature detected by the coolant temperature sensor SN2 is increased,in other words, as the in-cylinder temperature is increased, the timingof the preceding ignition is advanced (see Graph (b) in FIG. 16).Furthermore, as the in-cylinder pressure at the IVC time point detectedby the in-cylinder pressure sensor SN3 is increased, the timing of thepreceding ignition is retarded (see Graph (c) in FIG. 16). Furthermore,as the air-fuel ratio of the air-fuel mixture estimated from thedetection values of the plural sensors SN3, SN4, SN9, and the like isincreased, in other words, the concentration of the fuel in thecombustion chamber 6 is reduced, the timing of the preceding ignition isretarded (see Graph (d) in FIG. 16). With these configurations, it ispossible to sufficiently modify the property of the fuel so as toaccelerate the CI combustion.

Here, it has been understood that, in regard to the relationship betweenthe limit timing of the ignition timing, with which the flamepropagation occurs, (the limit timing: such limit timing that the flamepropagation occurs when the ignition timing is retarded from this limittiming) and the opening degree of the swirl valve 18, a change rate ofthe limit timing with respect to the opening degree of the swirl valve18 is lower in the range where the opening degree of the swirl valve 18is small than in the range where the opening degree of the swirl valve18 is large. By reflecting such a circumstance, in the above embodiment,as illustrated in Graph (a) in FIG. 16, the change rate of the timing ofthe preceding ignition with respect to the opening degree of the swirlvalve 18 is set to be lower in the range where the opening degree of theswirl valve 18 is small than in the range where the opening degree ofthe swirl valve 18 is large. In this way, it is possible to apply theenergy to the air-fuel mixture by the preceding ignition at theappropriate timing that confirms to a change characteristic of the limittiming as described above, and it is thus possible to sufficientlymodify the property of the fuel so as to accelerate the CI combustion.

For the similar reason, in the above embodiment, the change rate of thetiming of the preceding ignition with respect to the engine coolanttemperature is set to be lower in the range where the engine coolanttemperature is low than in the range where the engine coolanttemperature is high (see Graph (b) in FIG. 16), the change rate of thetiming of the preceding ignition with respect to the in-cylinderpressure at the IVC time point is set to be lower in the range where thein-cylinder pressure is high than in the range where the in-cylinderpressure is low (see Graph (c) in FIG. 16), and the change rate of thetiming of the preceding ignition with respect to the air-fuel ratio(A/F) is set to be lower in the range where the air-fuel ratio is highthan in the range where the air-fuel ratio is low (see Graph (d) in FIG.16). These characteristics are set on the basis of the characteristicsthat are studied in advance about the changes in the limit timing withrespect to the in-cylinder temperature, the in-cylinder pressure, andthe fuel concentration. By adjusting the timing of the precedingignition according to the characteristics, it is possible tosufficiently modify the property of the fuel so as to accelerate the CIcombustion.

In the above embodiment, the timing of the preceding ignition is set inthe early period or the middle period of the compression stroke.Accordingly, the preceding ignition can be carried out at timing afterthe fuel injection in the intake stroke and at which the piston 5 issufficiently advanced from the compression top dead center. In this way,it is possible to reliably modify the property of the fuel after theinjection by the preceding ignition so as to accelerate the CIcombustion while the unintended occurrence of the flame propagation bythe preceding ignition is avoided.

In the above embodiment, during the operation in the first operationrange A1, the swirl valve 18 is closed to such an opening degree thatthe swirl ratio of 1.5 or higher is secured. Accordingly, it is possibleto spread the intermediate product, which is produced by the precedingignition, over the large area of the combustion chamber 6 in the shorttime by the swirl flow. Then, by using this spread intermediate product,it is possible to effectively increase the combustion velocity in the CIcombustion that is simultaneously and frequently initiated in variousportions of the combustion chamber 6.

In the above embodiment, the main ignition timing by the ignition plug16 is adjusted such that the SI rate, which is the ratio of the heatgeneration amount by the SI combustion to the total heat generationamount in one cycle, matches the target SI rate, which is predeterminedaccording to the engine operation condition, during the SPCCI combustion(during the operation in the first and second operation ranges A1, A2).Thus, it is possible to increase the ratio of the CI combustion (thatis, to reduce the SI rate) as much as possible within the range wherethe combustion noise does not become excessive, for example. Togetherwith the effect exerted by the modification of the property of the fuelby the preceding ignition (the acceleration of the CI combustion), thiscan lead to as much improvement in the thermal efficiency as possible bythe SPCCI combustion.

Here, the preceding ignition, which is carried out prior to the mainignition, only functions to produce the intermediate product containingOH radicals (and thereby increase the combustion velocity in the CIcombustion). Thus, even in the case where the energy or the timing ofthe preceding ignition is changed, the SI rate or the initiation timing(θci) of the CI combustion is not particularly influenced by such achange. This means that the main ignition timing for obtaining thetarget SI rate can uniquely be specified independently from the energyor the timing of the preceding ignition. That is, according to the aboveembodiment, it is possible to specify the main ignition timing forobtaining the target SI rate with a high degree of accuracy whilecarrying out the preceding ignition to produce the sufficient amount ofthe intermediate product.

In the above embodiment, the operation range where the precedingignition and the main ignition are carried out is limited only to theportion on the low-load side in the range where the SPCCI combustion iscarried out (the first and second operation ranges A1, A2), that is, thefirst operation range A1, and the preceding ignition is not carried outin the second operation range A2 on the high-load side. Thus, it ispossible to effectively avoid abnormal combustion caused by the CIcombustion that is accelerated by the preceding ignition. That is, inthe case where the preceding ignition is carried out to produce theintermediate product such as OH radicals in the second operation rangeA2 on the high-load side, the combustion velocity in the CI combustionis excessively increased, which increases a possibility of the abnormalcombustion such as knocking. Meanwhile, in the above embodiment, thepreceding ignition is prohibited in the second operation range A2 on thehigh-load side. Thus, it is possible to effectively avoid the abnormalcombustion such as knocking or the like.

Even in the case where the preceding ignition is carried out in thesecond operation range A2, the above abnormal combustion can be avoidedby retarding the main ignition timing, for example. However, theretardation of the main ignition leads to the reduction in the torque,and the fuel injection amount has to be increased to compensate for thisreduction. As a result, the effect by the preceding ignition (a fuelconsumption improvement effect by the acceleration of the CI combustion)is canceled. Just as described, the prohibition of the precedingignition in the second operation range A2 has the equivalent resultwhile the control is simplified. Thus, it is desired to prohibit thepreceding ignition in the second operation range A2.

In the above embodiment, at least during the operation in the firstoperation range A1 in the warm period of the engine, the control isexecuted to carry out the SPCCI combustion while creating the lean A/Fenvironment where the air-fuel ratio (A/F), which is the ratio betweenthe air and the fuel in the combustion chamber 6, becomes higher thanthe stoichiometric air-fuel ratio. Accordingly, while the CI combustionis accelerated by the preceding ignition (and the thermal efficiency isthereby improved), the combustion temperature of the air-fuel mixturecan be kept low. Thus, it is possible to effectively suppress an amountof NOx produced in association with the combustion.

In the above embodiment, in the first operation range A1 where thepreceding ignition and the main ignition are carried out, the injector15 injects the fuel by dividing the fuel into two for the firstinjection and the second injection at the earlier timing than thepreceding ignition (in the intake stroke). Accordingly, when theinjection amount/the injection timing of the fuel by the first injectionand the second injection are set according to the engine operationcondition, it is possible to adjust the degree of the stratification (ora degree of homogeneity) of the air-fuel mixture in the combustionchamber 6 such that the appropriate SPCCI combustion is carried outunder each of the operation conditions.

For example, at the operation point P1 on the low-load side in the firstoperation range A1, the timing of the second injection is relativelyretarded (see FIG. 6(a)). At the operation point P2 on the high-loadside in the first operation range A1, the timing of the second injectionis relatively advanced (see FIG. 6(b)). Thus, it is possible to producethe appropriate air-fuel mixture in the combustion chamber 6 inconsideration of both of the ignitability and an emission property. Thatis, the timing of the second injection is retarded under the conditionthat the load is low and the fuel injection amount in one cycle issmall. Accordingly, the air-fuel mixture can be stratified such that theconcentration of the fuel in the central portion of the combustionchamber 6 is increased, and thus it is possible to improve theignitability on the low-load side. In addition, the timing of the secondinjection is advanced under the condition that the load is high and thefuel injection amount is large. Accordingly, the air-fuel mixture can behomogenized such that the excessively rich air-fuel mixture is notproduced locally, and thus it is possible to obtain the favorableemission performance.

However, in the first operation range A1, the fuel injection amount bythe first injection is set to be larger than the fuel injection amountby the second injection regardless of the amount of the load (at any ofthe operation points P1, P2). In this way, the fuel is not excessivelystratified, and thus it is possible to secure the favorable emissionperformance.

(10) MODIFIED EXAMPLES

For example, as illustrated in FIG. 16, the specified calculationformula having, as the input elements, the opening degree of the swirlvalve 18, the engine coolant temperature, the in-cylinder pressure atthe IVC time point, and the air-fuel ratio (A/F) of the air-fuel mixturein the combustion chamber 6 and having, as the output element, thetiming of the preceding ignition is used to determine the timing of thepreceding ignition. However, the method for determining the timing ofthe preceding ignition is not limited thereto.

For example, the timing of the preceding ignition may be determined fromthe above four parameters by using a map that is defined in advance. Indetail, it is considered to determine the timing of the precedingignition by determining basic preceding ignition timing on the basis ofa specified map that defines a relationship between some parameters ofthe above four parameters and a basic value of the timing of thepreceding ignition (the basic preceding ignition timing), by correctingthe determined basic preceding ignition timing according to the value ofthe other parameter(s), and the like. In addition, elements thatdetermine the timing of the preceding ignition are not limited to theabove four parameters (the swirl opening degree, the engine coolanttemperature, the in-cylinder pressure, and the air-fuel ratio), but alsothe engine load (the requested torque) and the engine speed may beconsidered as the parameters.

In the above embodiment, the swirl valve 18 is provided in one (thesecond intake port 9B) of the two intake ports 9A, 9B that are providedfor the single cylinder 2, and the opening degree of this swirl valve 18is increased/reduced to adjust the strength of swirl flow (the swirlratio). However, the method for adjusting the strength of the swirl flowis not limited thereto. For example, it is also possible to adjust thestrength of swirl flow by making a lift amount of the intake valve 11that opens/closes the first intake port 9A and a lift amount of theintake valve 11 that opens/closes the second intake port 9B differ fromeach other or by making opening/closing timing of these two intakevalves 11 differ from each other.

In the above embodiment, the preceding ignition is carried out in theearly period or the middle period of the compression stroke after thefuel injection by the first injection and the second injection iscompleted. However, the timing of the preceding ignition only needs tobe timing at which the fuel exists in the combustion chamber 6. Forexample, the preceding ignition may be carried out during the intakestroke. Furthermore, the number of the preceding ignition is not limitedto one in one cycle, and may be increased to two or more.

For example, as illustrated in Chart (a) in FIG. 19, the precedingignition may be carried out twice from the early period to the middleperiod of the compression stroke after the fuel injection by the firstinjection and the second injection is completed. Alternatively, asillustrated in Chart (b) in FIG. 19, the first preceding ignition may becarried out in a period between the first injection and the secondinjection, and the second preceding ignition may be carried out afterthe second injection is completed. At this time, the ignition timing ofeach of the preceding ignition only needs to be set according to thein-cylinder temperature, the in-cylinder pressure, the concentration ofthe fuel in the combustion chamber 6, and the flow rate of the gas inthe combustion chamber 6 as described above.

However, it is desired to set the number of the preceding ignition to beequal to or smaller than three. This will be described with reference toFIG. 20. FIG. 20 is a graph illustrating a relationship between thenumber of the preceding ignition and improvement allowance of a fuelconsumption rate (g/kWh). As illustrated in this drawing, the fuelconsumption rate is sufficiently improved when the preceding ignition iscarried out once. Meanwhile, the fuel consumption rate is gradually andfurther improved when the number of the preceding ignition is increasedto two or three. However, when the number of the preceding ignition isincreased from three to four, a value of the fuel consumption ratesubstantially remains the same. Just as described, since any effect canhardly be obtained by increasing the number of the preceding ignition tofour or more, it is desired to set the number of the preceding ignitionto three or less.

In the above embodiment, when the engine is in the warm state and isoperated in the first operation range A1 where the load and the speedare low, the control is executed to subject the air-fuel mixture to theSPCCI combustion while creating the lean A/F environment where theair-fuel ratio (A/F), which is the weight ratio between the air and thefuel in the combustion chamber 6, becomes higher than the stoichiometricair-fuel ratio (14.7), in detail, the environment where the air-fuelratio exceeds 20 and is lower than 35. However, the in-cylinderenvironment at the time when the SPCCI combustion is carried out in thefirst operation range A1 is not limited thereto. For example, inaddition to an amount of the air (the fresh air) corresponding to thestoichiometric air-fuel ratio, the EGR gas may be introduced into thecombustion chamber 6, so as to create lean G/F environment where theair-fuel ratio (A/F) substantially matches the stoichiometric air-fuelratio and a gas air-fuel ratio (G/F) as the weight ratio between theentire gas and the fuel in the combustion chamber 6 becomes higher thanthe stoichiometric air-fuel ratio, and the SPCCI combustion may becarried out in such a state. A value of the gas air-fuel ratio (G/F) inthe case where the SPCCI combustion is carried out under the lean G/Fenvironment just as described is preferably higher than 18 and lowerthan 50. Alternatively, in the case where the SPCCI combustion iscarried out in the same first operation range A1, one of two types ofthe environment may be adopted according to a temperature condition suchthat the lean A/F environment is created in the engine warm period inwhich the ignitability is easily secured and that the lean G/Fenvironment is created under a lower temperature condition than theabove (for example, in a semi-warm period).

In the above embodiment, the SI rate, which is the ratio of the heatgeneration amount by the SI combustion to the total heat generationamount by the SPCCI combustion, is defined as R1/(R1+R2) by using theareas R1, R2 in the combustion waveform illustrated in FIG. 7, and themain ignition timing is adjusted such that this SI rate matches thepredetermined target SI rate. However, various other methods fordefining the SI rate are considered.

For example, the SI rate=R1/R2 may be used. Furthermore, the SI rate maybe defined by using Δθ1, Δθ2 illustrated in FIG. 21. That is, in thecase where the crank angle period of the SI combustion (a combustionperiod on the advanced side of the inflection point X2) is set as Δθ1,and the crank angle period of the CI combustion (a combustion period onthe retarded side of the inflection point X2) is set as Δθ2, the SIrate=Δθ1/(Δθ1+Δθ2) or the SI rate=Δθ1/Δθ2 may be adopted. Alternatively,in the case where a peak of the heat generation rate in the SIcombustion is set as ΔH1, and a peak of the heat generation rate in theCI combustion is set as ΔH2, the SI rate=ΔH1/(ΔH1+ΔH2) or the SIrate=ΔH1/ΔH2 may be adopted.

In the above embodiment, the description has been made on the case wherethe energy of the preceding ignition is set according to the enginespeed and the engine load. However, the energy of the preceding ignitionmay be set to be constant regardless of the engine operation condition,and the high-temperature portion at equal to or higher than 850 K andlower than 1140 K may be generated around the spark only by adjustingthe timing of the preceding ignition.

Summary of Embodiment

The embodiment will be summarized as follows.

The control apparatus according to the embodiment is the controlapparatus for controlling the compression-ignition type engine thatincludes: the cylinder; the injector that injects the fuel into thecylinder; and the ignition plug that ignites the air-fuel mixture, inwhich the fuel injected from the injector and air are mixed, and thatcan carry out the partial compression ignition combustion to subjectsome of the air-fuel mixture to the SI combustion by the spark ignitionusing the ignition plug and to subject the rest of the air-fuel mixtureto the CI combustion by the self-ignition. The control apparatusincludes: the swirl generation section that generates the swirl flow inthe cylinder; the injection control section that controls the fuelinjection operation by the injector; and the ignition control sectionthat controls the ignition operation by the ignition plug. When thepartial compression ignition combustion is carried out, the ignitioncontrol section causes the ignition plug to carry out: the main ignitionin which the spark is generated in the late period of the compressionstroke or the initial period of the expansion stroke to initiate the SIcombustion; and the preceding ignition in which the spark is generatedat the earlier timing than the main ignition. When the partialcompression ignition combustion is carried out, the injection controlsection causes the injector to inject the fuel at such timing that thefuel exists in the cylinder at the earlier time point than the precedingignition. The timing of the preceding ignition is set to be moreadvanced when the swirl flow generated by the swirl generation sectionis gentle than when the swirl flow is strong.

According to such a configuration, when the partial compression ignitioncombustion is carried out, the main ignition, in which the spark isgenerated in the late period of the compression stroke or the initialperiod of the expansion stroke (that is, timing to generate the torque)to initiate the SI combustion, and the preceding ignition, in which thespark is generated at such timing that the fuel exists in the cylinderand that is earlier than the main ignition. Accordingly, when thetemperature around the spark (the arc) is increased to the appropriatetemperature by using this preceding ignition, it is possible to modifythe property of the fuel such that the thermal efficiency during the CIcombustion is increased while the flame propagation of the air-fuelmixture is suppressed. In detail, the fuel component (hydrocarbons) iscleaved by increasing the temperature thereof, so as to produce hydrogenperoxide (H₂O₂) and formaldehyde (CH₂O), and OH radicals from thesecomponents can be produced. OH radicals have the strong oxidationbehavior and are high in reactivity. Thus, when the intermediate productcontaining such OH radicals is produced in the cylinder after thepreceding ignition, the combustion velocity of the CI combustion as thephenomenon in which the spontaneous chemical reaction of the fuelcomponent occurs can be accelerated, which can improve the thermalefficiency.

Here, in order to modify the property of the fuel by the precedingignition as described above (to produce the intermediate product such asOH radicals while sufficiently suppressing the flame propagation), it isnecessary to appropriately adjust the timing of the preceding ignitionaccording to the various types of the environment in the cylinder. Theinventors of the present application performed the earnest study on thispoint and understood that the flame propagation was less likely to occurdue to the increase in the flow in the cylinder at the time of theignition when the ignition timing was advanced, and that the limittiming of the ignition timing, at which the flame propagation occurred,(the limit timing: such ignition timing that the flame propagationoccurred when the ignition timing was retarded from this limit timing)was retarded with the increase in the swirl flow, in other words, themost retarded timing in the range of the ignition timing, in which theflame propagation did not occur, was advanced with weakening of theswirl flow (however, the parameters other than the swirl flow wereconstant). That is, it was understood that, as the swirl flow wasgentle, the flame propagation was likely to occur due to weakening ofthe flow in the cylinder. Thus, in the case where the swirl flow wasgentle, the ignition timing had to be further advanced, and the ignitionhad to be carried out at such timing that the significant flow occurredin the cylinder. Otherwise, the flame grew without being blown off, andthe flame propagation occurred.

Meanwhile, in the above configuration, the timing of the precedingignition is more advanced when the swirl flow is gentle than when theswirl flow is strong. Accordingly, when the swirl flow is gentle and theflame propagation is likely to occur, the ignition timing is advanced.When the swirl flow is strong and the flame propagation is less likelyto occur, the ignition timing is retarded. In this way, it is possibleto apply the energy to the air-fuel mixture by the preceding ignition atthe appropriate timing that corresponds to the difference in the flamepropagation property. As a result, within the range where the flamepropagation of the air-fuel mixture does not occur or hardly occurs, itis possible to appropriately heat the air-fuel mixture to thetemperature zone in which the property of the air-fuel mixture issufficiently modified, and it is thus possible to reliably produce theintermediate product that contributes to the acceleration of the CIcombustion.

In order to appropriately modify the property of the fuel by thepreceding ignition, it is preferred to set the energy of the precedingignition to be lower than the energy of the main ignition. In detail,the energy and the timing of the preceding ignition are preferably setto such energy and such timing that the high-temperature portion atequal to or higher than 850 K and lower than 1140 K is generated aroundthe spark generated from the ignition plug and that the flamepropagation of the air-fuel mixture does not occur.

With these configurations, while the sufficient amount of theintermediate product is produced by the preceding ignition, it ispossible to avoid the intermediate product from being consumed prior tothe main ignition (by the unintended flame propagation by the precedingignition).

The swirl generation section can be the openable/closable swirl valvethat is provided in the intake port communicating with the cylinder. Inthis case, the ignition control section preferably controls the ignitionplug such that the timing of the preceding ignition is advanced with theincrease in the opening degree of the swirl valve.

With this configuration, the timing of the preceding ignition isadvanced with the increase in the opening degree of the swirl valve andthus with the reduction in the strength of the swirl flow. As a result,it is possible to appropriately modify the property of the fuel andaccelerate the CI combustion within the range where the flamepropagation does not substantially occur.

Here, it has been understood that, as the further detailed knowledge onthe relationship between the limit timing of the timing of the precedingignition, at which the flame propagation occurs, (hereinafter alsosimply referred to as the limit timing) and the opening degree of theswirl valve, the change rate of the limit timing with respect to theopening degree of the swirl valve is lower in the range where theopening degree of the swirl valve is small than in the range where theopening degree of the swirl valve is large. Thus, the ignition controlsection preferably sets the change rate of the timing of the precedingignition with respect to the opening degree of the swirl valve such thatthe change rate is lower in the range where the opening degree of theswirl valve is small than in the range where the opening degree of theswirl valve is large.

With such a configuration, it is possible to apply the appropriateenergy that conforms to the change characteristic of the limit timing asdescribed above by the preceding ignition, and it is thus possible tosufficiently modify the property of the fuel so as to accelerate the CIcombustion.

The timing of the preceding ignition is preferably set during the intakestroke or in the early period or the middle period of the compressionstroke.

In the case where the preceding ignition is carried out at the timingthat is sufficiently advanced from the compression top dead center, itis possible to reliably modify the property of the fuel after theinjection by the preceding ignition so as to accelerate the CIcombustion while the unintended occurrence of the flame propagation bythe preceding ignition is avoided.

In the case where the preceding ignition is carried out during theintake stroke or in the early period or the middle period of thecompression stroke, the sufficient interval is likely to be securedbetween the preceding ignition and the main ignition. Accordingly, thepreceding ignition and the main ignition are preferably carried out bythe single ignition plug having the single ignition circuit percylinder.

With such a configuration, it is possible to carry out the precedingignition and the main ignition by a simple method using the existingignition plug.

The number of the preceding ignition is not limited to one in one cycle,and the plural times of the preceding ignition may be carried out.However, according to the study by the inventors of the presentapplication, even in the case where the preceding ignition is carriedout more than three times, the obtained effect is hardly changed fromthe effect obtained when the preceding ignition is carried out threetimes. Thus, it is preferred that the number of the preceding ignitionin one cycle is equal to or less than three.

With such a configuration, it is possible to suppress wear of anelectrode of the ignition plug while securing the effect of thepreceding ignition to accelerate the CI combustion.

Preferably, the swirl generation section is the openable/closable swirlvalve that is provided in the intake port communicating with thecylinder, and the opening degree of the swirl valve is controlled suchthat, in the operating range where the preceding ignition and the mainignition are carried out, the swirl ratio in the cylinder becomes equalto or higher than 1.5.

Just as described, in the case where the relatively intense swirl flowwith the swirl ratio of equal to or higher than 1.5 is generated in theoperating range where the preceding ignition and the main ignition arecarried out, the intermediate product that is produced by the precedingignition can be spread over the large area of the cylinder in the shorttime by the swirl flow. Thus, by using this spread intermediate product,it is possible to effectively increase the combustion velocity in the CIcombustion that is simultaneously and frequently initiated in variousportions of the cylinder.

Preferably, the setting section for setting the target SI rate, which isthe target value of the ratio of the heat generation amount by the SIcombustion to the total heat generation amount in one cycle, accordingto the engine operation condition is further provided, and the ignitioncontrol section sets the main ignition timing on the basis of the targetSI rate set by the setting section.

In the case where the main ignition timing is adjusted to carry out thepartial compression ignition combustion that conforms to the target SIrate just as described, it is possible to increase the ratio of the CIcombustion (that is, reduce the SI rate) as much as possible within therange where the combustion noise does not become excessive, for example.Together with the effect exerted by the modification of the property ofthe fuel by the preceding ignition (the acceleration of the CIcombustion), this can lead to as much improvement in the thermalefficiency as possible by the partial compression ignition combustion.

Here, the preceding ignition, which is carried out prior to the mainignition, only functions to produce the intermediate product containingOH radicals (and thereby increase the combustion velocity in the CIcombustion). Thus, even in the case where the energy or the timing ofthe preceding ignition is changed, the SI rate is not particularlyinfluenced by such a change. This means that the main ignition timingfor obtaining the target SI rate can uniquely be specified independentlyfrom the energy or the timing of the preceding ignition. That is,according to the above configuration, it is possible to specify the mainignition timing for obtaining the target SI rate with the high degree ofaccuracy while carrying out the preceding ignition to produce thesufficient amount of the intermediate product.

Preferably, the ignition control section executes the preceding ignitionand the main ignition only in the portion on the low-load side of theoperation range where the partial compression ignition combustion iscarried out.

With such a configuration, it is possible to effectively avoid theabnormal combustion caused by the CI combustion that is accelerated bythe preceding ignition. That is, in the case where the precedingignition and the main ignition are carried out uniformly in the rangewhere the partial compression ignition combustion is carried out, thecombustion velocity in the CI combustion is excessively increased on thehigh-load side in the same range, which increases the possibility of theabnormal combustion such as knocking. Meanwhile, in the aboveconfiguration, the preceding ignition is carried out only in the portionon the low-load side of the range where the partial compression ignitioncombustion is carried out, and the preceding ignition is prohibited onthe high-load side. Thus, it is possible to effectively avoid theabnormal combustion such as knocking.

Preferably, in the operation range where the preceding ignition and themain ignition are carried out, the partial compression ignitioncombustion is carried out under the lean A/F environment where theair-fuel ratio as the ratio between the air and the fuel in the cylinderexceeds 20 and is lower than 35, or the partial compression ignitioncombustion is carried out under the lean G/F environment where the gasair-fuel ratio as the ratio between the entire gas and the fuel in thecylinder exceeds 18 and is lower than 50 and the air-fuel ratiosubstantially matches the stoichiometric air-fuel ratio.

With such a configuration, it is possible to suppress the combustiontemperature of the air-fuel mixture to be kept low while the CIcombustion is accelerated by the preceding ignition (and the thermalefficiency is thereby improved). Thus, it is possible to effectivelysuppress the amount of NOx produced in association with the combustion.

Preferably, the injection control section causes the injector to carryout the second injection, in which the fuel is injected prior to thepreceding ignition, and the first injection, in which the fuel isinjected prior to the second injection.

With such a configuration, when the injection amount/the injectiontiming of the fuel by the first injection and the second injection areset according to the engine operation condition, it is possible toadjust the degree of the stratification (or the degree of thehomogeneity) of the air-fuel mixture such that the appropriate partialcompression ignition combustion is carried out under each of theoperation conditions.

In the above configuration, further preferably, the injection controlsection controls the injector such that the fuel injection amount by thefirst injection is larger than the fuel injection amount by the secondinjection.

In the case where the injection amount by the first injection at theearly injection timing is relatively increased just as described, it ispossible to effectively avoid the excessive stratification of the fueland degradation of the emission performance.

The control apparatus in the embodiment can also be said to have thefollowing characteristics. More specifically, the control apparatus isthe apparatus for controlling the engine that includes: the cylinder;and the injector and the ignition plug disposed in the manner to facethe cylinder, and includes: the swirl adjustment device that adjusts theswirl flow generated in the cylinder; the temperature adjustment devicethat adjusts the in-cylinder temperature as the temperature in thecylinder; and the controller that includes the electric circuitelectrically connected to the injector, the ignition plug, the swirladjustment device, and the temperature adjustment device and thatoutputs the control signal to each of the injector, the ignition plug,the swirl adjustment device, and the temperature adjustment device. Thecontroller has: the injection control section that drives the injectorto inject the fuel at the specified timing corresponding to theoperation state of the engine; the in-cylinder temperature adjustmentsection that drives the in-cylinder temperature adjustment device toadjust the in-cylinder temperature such that the air-fuel mixture of thefuel injected from the injector and the air is subjected to the flamepropagation combustion by the spark ignition using the ignition plug andthat the compression self-ignition combustion is carried out after theinitiation of this flame propagation combustion; the first ignitioncontrol section that drives the ignition plug for the spark ignitionafter the fuel injection by the injector; the second ignition controlsection that drives the ignition plug for the spark ignition after thespark ignition by the first ignition control section, so as to subjectthe air-fuel mixture to the flame propagation by the spark ignition; andthe ignition timing setting section that sets the timing of the sparkignition by the first ignition control section to be more advanced whenthe swirl flow is gentle than when the swirl flow is strong on the basisof the control amount of the swirl adjustment device.

With such a configuration, similar to the above-described configuration,it is possible to modify the property of the fuel in the manner toincrease the combustion velocity of the combustion by the compressionself-ignition, and it is thus possible to improve the thermalefficiency.

1. An apparatus for controlling a compression-ignition type engine thatincludes: a cylinder; an injector that injects fuel into the cylinder;and an ignition plug that ignites air-fuel mixture, in which fuelinjected from injector and air are mixed, and that can carry out partialcompression ignition combustion to subject some of the air-fuel mixtureto SI combustion by spark ignition using the ignition plug, and subjectthe rest of the air-fuel mixture to CI combustion by self-ignition, thecontrol apparatus for the compression-ignition type engine comprising: aswirl generation section that generates a swirl flow in the cylinder; aninjection control section that controls fuel injection operation by theinjector; and an ignition control section that controls ignitionoperation by the ignition plug, wherein, when the partial compressionignition combustion is carried out, the ignition control section causesthe ignition plug to carry out: main ignition in which a spark isgenerated in a late period of a compression stroke or an initial periodof an expansion stroke to initiate the SI combustion; and precedingignition in which the spark is generated at earlier timing than the mainignition, wherein, when the partial compression ignition combustion iscarried out, the injection control section causes the injector to injectthe fuel at such timing that the fuel exists in the cylinder at anearlier time point than the preceding ignition, and wherein timing ofthe preceding ignition is set to be more advanced when the swirl flowgenerated by the swirl generation section is gentle than when the swirlflow is strong.
 2. The control apparatus for the compression-ignitiontype engine according to claim 1, wherein energy of the precedingignition is lower than energy of the main ignition.
 3. The controlapparatus for the compression-ignition type engine according to claim 2,wherein the energy and the timing of the preceding ignition are set tosuch energy and such timing that a high-temperature portion at equal toor higher than 850 K and lower than 1140 K is generated around the sparkgenerated from the ignition plug and that flame propagation of theair-fuel mixture does not occur.
 4. The control apparatus for thecompression-ignition type engine according to claim 1, wherein the swirlgeneration section is the openable/closable swirl valve that is providedin the intake port communicating with the cylinder, and wherein theignition control section controls the ignition plug such that the timingof the preceding ignition is advanced with an increase in an openingdegree of the swirl valve.
 5. The control apparatus for thecompression-ignition type engine according to claim 4, wherein theignition control section sets a change rate of the timing of thepreceding ignition with respect to the opening degree of the swirl valvesuch that the change rate is lower in a range where the opening degreeof the swirl valve is small than in a range where the opening degree ofthe swirl valve is large.
 6. The control apparatus for thecompression-ignition type engine according to claim 1, wherein thetiming of the preceding ignition is set during an intake stroke or in anearly period or a middle period of the compression stroke.
 7. Thecontrol apparatus for the compression-ignition type engine according toclaim 6, wherein the preceding ignition and the main ignition arecarried out by the single ignition plug having a single ignition circuitfor each of the cylinders.
 8. The control apparatus for thecompression-ignition type engine according to claim 1, wherein thenumber of the preceding ignition in one cycle is set to three or less.9. The control apparatus for the compression-ignition type engineaccording to claim 1, wherein the swirl generation section is theopenable/closable swirl valve that is provided in the intake portcommunicating with the cylinder, and wherein an opening degree of theswirl valve is controlled such that a swirl ratio in the cylinderbecomes equal to or higher than 1.5 in an operation range where thepreceding ignition and the main ignition are carried out.
 10. Thecontrol apparatus for the compression-ignition type engine according toclaim 1 further comprising: a setting section that sets a target SI rateas a target value of a ratio of a heat generation amount by the SIcombustion to a total heat generation amount in one cycle according toan engine operation condition, wherein the ignition control section setstiming of the main ignition on the basis of the target SI rate that isset by the setting section.
 11. The control apparatus for thecompression-ignition type engine according to claim 1, wherein theignition control section carries out the preceding ignition and the mainignition only in a portion on a low-load side of the operation rangewhere the partial compression ignition combustion is carried out. 12.The control apparatus for the compression-ignition type engine accordingto claim 1, wherein, in the operation range where the preceding ignitionand the main ignition are carried out, the partial compression ignitioncombustion is carried out under lean A/F environment where an air-fuelratio as a ratio between the air and the fuel in the cylinder exceeds 20and is lower than 35, or the partial compression ignition combustion iscarried out under lean G/F environment where a gas air-fuel ratio as aratio between entire gas and the fuel in the cylinder exceeds 18 and islower than 50 and the air-fuel ratio substantially matches astoichiometric air-fuel ratio.
 13. The control apparatus for thecompression-ignition type engine according to claim 1, wherein theinjection control section causes the injector to carry out secondinjection, in which the fuel is injected prior to the precedingignition, and first injection, in which the fuel is injected prior tosaid second injection.
 14. The control apparatus for thecompression-ignition type engine according to claim 13, wherein theinjection control section controls the injector such that a fuelinjection amount by the first injection is larger than a fuel injectionamount by the second injection.
 15. An apparatus for controlling anengine that includes: a cylinder; and an injector and an ignition plugdisposed in a manner to face the cylinder, the control apparatus for acompression-ignition type engine comprising: a swirl adjustment devicethat adjusts a swirl flow generated in the cylinder; a temperatureadjustment device that adjusts an in-cylinder temperature as atemperature in the cylinder; and a controller that includes an electriccircuit electrically connected to the injector, the ignition plug, theswirl adjustment device, and the temperature adjustment device and thatoutputs a control signal to each of said injector, the ignition plug,the swirl adjustment device, and the temperature adjustment device,wherein the controller has: an injection control section that drives theinjector at specified timing corresponding to an operation state of theengine so as to inject fuel; an in-cylinder temperature adjustmentsection that drives the in-cylinder temperature adjustment device toadjust the in-cylinder temperature such that air-fuel mixture of thefuel injected by the injector and air is subjected to flame propagationcombustion by spark ignition using the ignition plug and thatcompression self-ignition combustion is carried out after initiation ofthis flame propagation combustion; a first ignition control section thatdrives the ignition plug for the spark ignition after the injectorinjects the fuel; a second ignition control section that drives theignition plug for the spark ignition after the spark ignition by thefirst ignition control section so as to subject the air-fuel mixture tothe flame propagation combustion by said spark ignition; and an ignitiontiming setting section that sets timing of the spark ignition by thefirst ignition control section to be more retarded when the swirl flowis gentle than when the swirl flow is strong on the basis of a controlamount of the swirl adjustment device.
 16. The control apparatus for thecompression-ignition type engine according to claim 5, wherein thetiming of the preceding ignition is set during an intake stroke or in anearly period or a middle period of the compression stroke.
 17. Thecontrol apparatus for the compression-ignition type engine according toclaim 5, wherein the swirl generation section is the openable/closableswirl valve that is provided in the intake port communicating with thecylinder, and wherein an opening degree of the swirl valve is controlledsuch that a swirl ratio in the cylinder becomes equal to or higher than1.5 in an operation range where the preceding ignition and the mainignition are carried out.
 18. The control apparatus for thecompression-ignition type engine according to claim 5 furthercomprising: a setting section that sets a target SI rate as a targetvalue of a ratio of a heat generation amount by the SI combustion to atotal heat generation amount in one cycle according to an engineoperation condition, wherein the ignition control section sets timing ofthe main ignition on the basis of the target SI rate that is set by thesetting section.
 19. The control apparatus for the compression-ignitiontype engine according to claim 5, wherein the ignition control sectioncarries out the preceding ignition and the main ignition only in aportion on a low-load side of the operation range where the partialcompression ignition combustion is carried out.
 20. The controlapparatus for the compression-ignition type engine according to claim 5,wherein, in the operation range where the preceding ignition and themain ignition are carried out, the partial compression ignitioncombustion is carried out under lean A/F environment where an air-fuelratio as a ratio between the air and the fuel in the cylinder exceeds 20and is lower than 35, or the partial compression ignition combustion iscarried out under lean G/F environment where a gas air-fuel ratio as aratio between entire gas and the fuel in the cylinder exceeds 18 and islower than 50 and the air-fuel ratio substantially matches astoichiometric air-fuel ratio.